





















































































































































































































































































































































-- ./ - 4 c? S -V| *j * 

4 'ft 

* 

CK ^ 


& " 




^ ; \° $* 

^^,#V o 
>*••'• <? .>••'. 






0 * 


^ V..™ .,, - 

r oo^ ^ r 

^ *+ - €S 

O’ 0 s o ’ A^' ■%. V 91 "'^ 

x > 1 . 1 * 0 , > <V -.,, *, C X v * 

L. . ** *r> ,-> 

^ l • ; ri®s : ^ ** ^ • 

* v^%- '. W-% » l • ./ V S . 

.* * . \yrl' a %.’'•■<' &' Ss r " 

^ ‘ * V? * 


**o 0 ^ 






v * o 

,0 'O. <^ 0^0 

N* c M , */> - N 

<V s '/, c* 

A®l * r>^v. •>, * 'x 

Rr> < 


0 

s (V 






p> ^ 




N G 




"% 

h -p --"Jji c 

-V * * 4 “ 

o o X 

* -n*. 

^ - 


^ o x : w, 

> p a : ,> /• 

♦ * *- V V *> * • ‘ '* *" H, 

\\ x ' '^p. 


v \, > 


A” V- 




p.- 

A * 


«■^‘ ; 'VV"«/o ) 

7- 




«\ sC> ! ✓ 

^ *>*♦>• ° ^ 0 x >' ^ s * o ^ *> " .O' s s 

; A', % A c;V * 

■ 1: ^ :{ ,:.•' » ; 

* * >* “ ^ ^ aV ■•?/• 



0 v c o^ «■ * ^/ l 

*> x ^ " 

oo x , ^ r \A * ^ 

* wf-"- -... -• / * , 0 o 

v # ^ * 77 ^' ^°' s * *, * a n ° * 

.> '°j > , 0 " s .w * cv 

r 7 ' •»> ^ ^ - -X *<■ a 

A 



• «A> rA 

* «/* - 0 %^. 

■ 

^ AV 

P V A' 

4* 

A. 

o 

» .0 *+ ' ’Sgwy; n“ 

^ v 7 , v ' o- 

V- V N > ‘ .0^.* ' s * • r 

* *Mfi 7 / ^ 


rj, ^ 

fl W c * ^ ^ ^ ,^-y 

. 0o «. 






'’.^L 

<s J v, 

is ^ *- 

v V • ^ » II I A ° 

V ^ 0 / > 

%. .* A° 

(>,**• *9 ■} ° 

; / 


^ > 6 t> > > 

r*. - ^ .a : . <» 

</* 0v^ 

* V''-'>V^*/V ** 1 a o; v 





N° 7 

c . ft o 

A- .'kv*-, '-^. . »’.*• 


+ '+.1/ * 

/ ^ -.-A 

, ♦»! \ "' 


: ^f f7 ' aV'Op c%'^w* 'l-7#, ; . ; • v77> J * % % 

* * y -A - V^V oV A- x 

. '-V/' A % ^ < O'"^’ .< - ^ ^o. ^ A 0 ^ Q N c C 

* X ^ N ,* v ' * \'^ o^ <.•*«*/% 44 °o * 0 C ^ 


\° q. 'V & 

r x • V '> ^ ^ y .> 

.0 -o * , N o ’ A 3 

N % v > ^ 

<“• 



!) * 



. =. - J .A 

V’'‘V.‘*JL**-> 


, ;> b o X 


>° x 7 '% * 

x #‘ ^ *■ 
V X ^ '■ * » A > 


VJ f \ % $WJ / 

N s" .A A ' „ ., ^ . 0 ^ 

A 4 > ’ » * l O Q> c 0 N 

x°®<. 


K°q<v 

'O Xi>s ~ >* o,^' V ^3t> ; , \» s 0 n "* 

O. V X * , \ 0 ' > A° C * 

o \$ : o ‘. .fc-- ^ '<7 


v * 

« A x ' * . v> ^ 

* V ■>+ °+ ^ * <A 

A A *o.v-* -CA 

#' % ,0^ 

A - iT i - >. . ^ A o . 

° O' 






7— 



^ •, ,*■!,# A 


f-f y .{. N 

s 

^ /“ > 

* »1 '* 

ft 

< /A O 

gA Z 



) 



H - "•' 



: \° * 

S 


V 

o 0 X 


c f... ..,V •'•’/*' -1 

- JgM| - ^ # * v:A, .. 

-,m* Wa^-v- 7 /% '•«« 

n i •' v \. -.:. ; V "f-f.M*. % f o^.»“ 

••* -®;. : - O0 < r .,-;J^' ^ v* - 

p ^ W" ^! i? ~ 7 ^ 

\f°\s- r,r'^** N ° ./a* o /V > ' i ' , /s' '^° > *"'* A / %' > 

*K<^ £r * * ^; r> v V ^ «T\S 5 >' />L ^ ^ ^ * ^• / 

■ -' 0 - 4 0 ,v..* 

6 rL AX L v '• + 




vr/' 4 - ^ ^ 

^ o ^ ^ ^ 





- ^ ^ 
« A > 


X 0o x. 



'<J 0 « 


o 5 ^ 









' 0 » v 




o o' 


° ^ITV' A / 

O /r/^sSLilLTl ^ 'p/ ^ 


1 • * 



r A V 


X°°^ 



v '* . *'’ .V' v 

X /V 

' ^ p* .- 


o * \ 




V 

.o v 

O vj 


«b & o 

K O 
^ *•» 

^ ^ " 


v ,.'*,To^ /’ % ‘■•Trx <? KWi c < 

. > vO X S -/// O V* * X * 0 / *> v 0 V ' '// 

[S4. ~1>.£ * : h^ a,.,** >a¥a; ■fe..«» 4 . ' ; i : 






X 0, x 





o o' 


4 '/, • - ^5 

^ > \V ^ * 

^ c*x ^ 



V ^ 



\ 


OO^ 



✓ A>, X: * * * 


* « 


o 0 

X X- 

y a O '/W.' «> X * 

, <?#.'* a no’ x \^ #o s .., %**h*' v* X. * 

X x* \> »'*°^ -> '* c- v »'>'">/ ^ 

- ^y :$&\\ ;rjli|‘- *V’-S'u 

X ** v x -Am* / X xS|x X\ „ 

.V 0 ..'«, X"-•'•'V\.,V 0 • ^./ ^>\ ■-> * Sj *' 1 ' .cA-«. ••% 

“ r! • -*- '- •?, .»<* **/$}*;. ° , 0 »'<s5Swx <«, .<♦ 


W 







S> fr 


w 



v- V 


x° X. 




\ 


o ©. 



- -p 



x V 


. ■ —J - 

v * rP 


\ <• 


,V--y,- Xvx>-/- ->y; ,V 

? A. r ' X C.P * 4 ^V. \Y' » A, »* X. C,P * * %■ A^’ * jfCvWA. o X 

m t <b s " ^ v .-• : v $> - mWM ° z <?*>■ r #^g« 




fr * 



& * 

y\ 

O ^ 

k \ s A > » 

v _ )'* , 

* •» ' X 

-* sr 

X- 



,V‘ 




> vf>, ' 1 \ 


\ I » 


•\° ^ 




nA -r 


’ -V X 


& » 

^ y /*\s" A X ,. %. y 0 .V- a O Nc 

vA ,i'»* c »v c 

© ^ ^ « a N ' c^ , '»V l l‘ S aO 0 '* j 4^' 

••'“ ^’om, \ *»N°' v v\,.., % 81 s s ** r ', r/ C' 

r^\\^ ^ : % 4 

•^ v v?<'W/./^ .** v% '* 


t * 0 


y o o 


X o N 



x 0 ^. 




o 


0 


> % \y<V"S a~> 

a-x> *">•;• 

X X" <v 
; - '/b c ^ v o 

.V ./> * 

AV > o 

» y >^ ^ 

'A V" ; *MZZ%' 





X 0o x. 







O, '# „ . ^ A, f , 

» V ’-• v'>' ••' 

"* '% t $‘ * >■ -\'» ^ 
^ V - ? )£Sl % /. •<* S 

//M ™ J - A% ; ^ J L ' - aV «p 



0^ (° NC ^ 

O «> / 'f* 

^ 'X 






« I A AsA s • » , ®/i. * 9 N ° ' . ^ 

X lV 



a V 




v . , ;...,v*- > /.- 

» % / * 
i - .v ^ 


« y .^ ^ 




a v ~ *. n A v ^ 

X' , <X ^ /,>««. s ^ v . . 'X " - l x 

0 ■ " X .# v° ^ * X Xj c p- . 

\° X * ° ^ 'cf*. 





,0o 





_ ^ / s S 7 

1 * * > a\ \ l fl a 

^ v V - / V'? y 2 -» ^ O 

A' 

; ^ « 


0 o X 


x°X 



Xo N 

C X -7 

«K 

o - ' * ■> - o ’ ^ 


v e . >,W.' X X "‘‘W,? '■ o e 'W.' A’ 0 o 

^ • •-/V* ’ ' ” \X ' • • , • 1 ’ 0 ^-X' • • ' X. ’ K ” v V A • » , © " ' / s VV < * -c. 

- x ,v *$'-r% r 0 >k, ^ !. V ..^* »MAo Xxx -• Sr® - 

A A zWZ^?; aV^ :W\¥* 

* ^ ^ % ~>s'£Ws & '* \y&rs ■* ^ 

-,%^'^ .%" ii 7..».<».7...V” /<•■*/ 

^ .t v ^/y97^ 1 . o C ♦ ^>v.^ y ^ 

✓ 
s 

$ % 0 r "o / % XW^'o - - c, .r 

' ^ .S' V * 9 * 0 % ./s' ’ ’ 'ft%. 5H ° ‘ V — 




'W 



: ^ 



>- \ 


V * o 




n a r 

>y " \ *" U 7itU3A\\\^ w a >, 

L» -<‘ 

v- . ^ v- *► s 

■fer^ s 4 a o y o» >. 4 >o <. ^,,v' X 

4 * a\ . v i « # "X, .o^e. a' <, 

X s AtoL* o 0° X, X ' 


’V, 


>- V 


X 00 ^. 



O ■ 





4* 

\0 J 

\ $ “ « 

^■ '■\ “.VJ 



y VMA-* {» 1 » 

O. ' * . « > .\V 


o o 
oS ^ 

NT 

















THE DESIGN AND CONSTRUCTION 

OF 

INTERNAL-COMBUSTION 

ENGINES 

A HANDBOOK FOR 

DESIGNERS and BUILDERS of GAS and OIL ENGINES 
By HUGO $ULDNER 

Chief Engineer and Director, Giildner-Motoren-gesellschaft, Aschaffenburg 

TRANSLATED FROM THE SECOND REVISED EDITION 

WITH ADDITIONS ON 

AMERICAN ENGINES 

BY 

H. DIEDERICHS 

Professor of Experimental Engineering 
Cornell University 

728 illustrations anb 36 JMfotng •■piatrs 


\ 

NEW YORK 

D. VAN NOSTRAND COMPANY 

23 MURRAY AND 27 WARREN STREETS 

1910 



A 


& 




Copyright, 1910 

BY 

D. VAN NOSTRAND COMPANY 


Entered at Stationers’ Hall Court, London, England 


All rights reserved 




lEhe Scientific $ress 
Jtobcrt Brummonb anb Comparro 
Nett' Hark 


©Cl. A 2 53 9 35 





AUTHOR’S PREFACE TO THE FIRST EDITION 


Germany’s gas engine industry justly enjoys an international reputation. Whether 
judged by the standards of age and experience, commercial importance or originality 
of design, it leads the world. Four-cycle and two-cycle engine, explosion and constant- 
pressure engine, automobile construction, and commercial utilization of blast-furnace gas 
—in short, everything that has served to lay the foundation of the industry and that 
has helped to make it vital and important is either the product of German thought, 
or was first practically realized on German soil. 

Our technical literature on the subject shows the same high degree of development. 
There are a number of important books as well as occasional articles treating of 
special fields of the industry, which for thoroughness and depth of scientific treatment 
hardly find their equal. But as far as the industry itself is concerned such treatises 
and publications are of little real and practical use. Writers are apt to say a good 
deal more about the general construction of existing internal-combustion engines than 
about the details of the problems involved, and as a matter of fact they generally get 
more information from the practical designer than they give to him. This objection 
holds against most of the technical handbooks now existing. Outside of the thermo¬ 
dynamic and thermal principles involved, at best so arranged as to serve a real purpose, 
such books very often contain nothing more than a series of descriptions of existing 
engines, illustrated by figures giving little idea of proportion of parts, and a few 
results of tests. Of how much real use are such works to the practical man in the 
field? Taking it altogether he usually finds what to him is ancient history, and 
whatever is new to him is often presented in such form as to make it unavailable to 
his needs. The trouble is that both text and illustration are arranged to suit the 
layman, and consequently the builder or designer finds little of practical value. Thus 
it happens that our present handbooks on internal-combustion engines have but little 
circulation among the men in the field, but are used mostly by those who first wish to 
study the nature of the internal-combustion engine and for that purpose require a 
more or less popularly written general treatise. 

The reasons for this condition of affairs are not far to seek. The gas engine 
originated and grew through experimentation and experience. All the improvements 
that were made and are still being made are the result of experimentation, and only 
in the course of the latter do the men actually engaged in the development of the 
industry acquire their special knowledge and experience. New developments in the 
industry are presented to outsiders usually in the perfected state, and if after such 
presentation scientific investigation takes hold of the subject, it—with very few 

iii 



IV 


AUTHOR’S PREFACE TO THE FIRST EDITION 


exceptions—concerns itself mainly with the theoretical thermodynamic facts and 
principles involved. To thus confine attention to a field, which is really only a small 
part of the engine constructor’s actual work, and which is in reality of much less 
importance than is commonly assigned to it, simply amounts to a failure on the part 
of scientific men to appreciate a real need. 

The author desires not to be misunderstood in making the above statements. No 
one doubts for a moment that thorough tests made on internal-combustion engines are 
of the highest scientific importance, besides possessing an indirect practical value, 
although the final results of such investigations are generally of a purely theoretical, 
and often even of hypothetical, nature and can consequently be practically utilized 
only in isolated instances. The criticisms made are not aimed at scientific investigation of 
the thermal problems involved, but at the highly exaggerated valuation put upon it as 
an auxiliary in gas engine design and construction, and at the narrow viewpoint which 
this tendency has succeeded in impressing upon all experimental work connected with 
the gas engine. The latter effect is strongly manifest in our technical literature as 
well as in our technical instruction. For some decades past science has apparently 
considered only the physical and thermal phenomena connected with the gas engine as 
worthy of its attention, and, in doing so, has seriously neglected the investigation of 
the static and kinetic features involved, that is, has paid little attention to the theory 
of the design of the internal-combustion engine. Complete volumes are filled with 
thermodynamic discussions, while a few pages will cover all that is written concerning 
constructive and technical details. 

There is little doubt that these things have constitiuted a serious drawback in the 
development of the industry and have worked considerable harm. For, even to-day, 
each manufacturing concern is compelled to determine for itself the best forms, the 
best mode of action and the allowable safe stresses for each essentially new detail of 
gas-engine construction, and is not able in any way to avail itself of the results of 
investigations made and of the experiences gained in probably a large number of 
similar cases in other works. As a consequence, many futile experiments are repeated 
again and again. The continued great loss of time and money incident to this proceed¬ 
ing could at least in large part be with certainty avoided if scientific investigation, 
instead of almost entirely confining itself to the thermodynamic side of the problem, 

could be induced to take up with equal energy the mechanical and kinetic side with 

special reference to the problems of actual design and practical construction. The 
solution of the entire problem does not rest entirely in the domain of thermodynamics; 
on the contrary, every technically trained man and every practically trained mechanic 
can do his share toward it. Under such conditions, technical literature, in so far as 
it pertains to the gas engine, would soon assume a quite different aspect and would 
attain a higher degree of importance in the practical field than it at present possesses. 

It is the aim of this book to make a beginning at clearing the ground along the 
lines thus defined and to be considered as the first modest contribution to this cause. 
Out of the fund of practical experience, gathered during the past fifteen years in 

positions of executive capacity, it has been written by a gas-engine designer and 
builder for others engaged in the same work, with the idea of presenting to the 

technical man, and especially to the gas-engine designer, a reliable handbook and to 
the gas-engine builder new in the field a practically useful guide. That there is real 
need for a book of this kind for both classes of men, the author believes to have 
shown in what has been said. 

Concerning the scope and manner of treatment of the material entering into the 


AUTHOR’S PREFACE TO THE FIRST EDITION 


v 


various problems, the book may speak for itself. The entire internal arrangement was 
in reality already fixed by the purpose for which the work is va’itten. 

The first part 1 of the book consists of a critical discussion of the most important 
of the older types of internal-combustion engines, to serves both as an introduction to 
the field in general and to give a brief resume concerning constructions already tried 
and in part abandoned. 

The second part treats first of the thermodynamics of the gas engine in a form 
and to an extent which the real purpose of the book seems to justify. This is 

followed by a critical examination of the various events of the gas-engine cycles, a 

subject which should be welcomed alike by the designing-room and the shop and 
which, on account of its importance, is quite likely to receive further elucidation and 
additions at the hands of others. 

The most extensive part of the book, however, the one which brings out most 
clearly the real aim of the entire work, is the third part, intended to serve as an 
everyday working guide to the designer and constructor. The author would like to 
look upon this as the beginning of a “science of gas-engine construction” whose aim 
it shall be to suitably extend the rules of general design and to adapt them to the 

special requirements of gas-engine construction, to bring together a selected collection 

of dependable data, and of approved forms of construction to serve as a reliable guide 
for new constructions, and thus save both designer and builder a great deal of 
unnecessary and costly trying-out and experimentation. This aim calls for the most 
extended use of drawings, true to detail, showing approved designs, and the more 
dimension figures are given in them, the better. The dimensions of machine parts 
tried and found satisfactory in actual service constitute statistics of peculiar value to 
the designer, and not infrequently such figures express rules of design not yet stated 
clearly in words. In fact, in cases where the underlying rules and laws are not yet 
clearly understood, such combinations of drawing and dimension find their most 
important usefulness. This viewpoint entirely determined the manner of illustrating 
the book, and especially of the third part of the same. If this practice should serve 
as an example followed by others, it can only prove of benefit to technical gas- 
engine literature, in general, because the hitherto much preferred method of banishing 
all dimension figures from text illustration has resulted in merely turning these in¬ 
dustrial documents into useless pictures. 

While the third part shows the method of designing the various separate machine 
parts, it is the main purpose of the fourth part to show how these various parts are 
utilized in combination in the modern types of internal-combustion engines and to give 
some information concerning the field of application, erection, etc., of the various forms. 
The results of the latest tests also reported in this part serve to show the present 
economic position of the gas engine and its standing as compared with other sources 
of power. 

The fifth part, treating of the gas-engine fuels and combustion in the gas engine, 
is also mainly arranged to suit the needs of the designer, but there is little doubt that 
the shopman will also find many things of service to him. Even engine builders 
grown gray in the business are sometimes not entirely clear on the various phenomena 
attending combustion in gas engines, nor how important all means tending to promote 
complete combustion really are and what means are available. It requires a study of 


1 Translator’s Note. The parts of the book have been rearranged, see Translator’s Preface. 



VI 


AUTHOR’S PREFACE TO THE FIRST EDITION 


the combustion process to see that, in a certain sense, air is power as far as the gas 

engine is concerned, and that to neglect this fact simply means a waste of fuel. 

In order that the designer and builder not familiar with thermodynamics may find 
the book entirely useful, a brief treatise of Thermodynamics and Thermochemistry is 
given in the Appendix, to avoid complicating the main part of the book with elemen¬ 
tary discussions. The rest of the subject matter of the Appendix is a miscellaneous 
collection of information on several topics, which, while it will probably not be reg¬ 
ularly used, will now and then prove of interest and value. 

Taking it altogether, this work is written “out of practice for practice;” it is 

nevetheless hoped that it will be of some use in technical instruction. For obvious 

reasons, the author, in making this statement, has in mind mainly the technical high 
school. The training which these schools at present impart to the engineering graduate 
along the line of gas-engine design amounts to next to nothing. Every chief designer 
can corroborate this statement from his own experience, no matter from where his 
assistants may come. That this condition of affairs should change is merely an 
industrial necessity and consequently a just demand on the part of the gas-engine 
industry. 

The objection to this statement of the case is the old one that “the high school 
should not be called upon to develop specialists.” How long this view of the question 
can be maintained in the face of the steady readjustment now at work in the industrial 
field is problematical. What is not open to question is the fact that the gas-engine 
industry has to-day attained- an extent and importance which make it barely secondary 
to steam-engine and machine-tool industry and to the electrical manufacturing interests. 
Now it is well known that the technical high schools in their present courses of study, 
examinations, and laboratory equipments, try to meet the practical requirements of 
these branches of our industry to the widest possible extent. It is therefore strongly 
in the interest of the gas-engine builder to see that those students who intend later to 
specialize in his branch of industrial activity, are given sufficient opportunity to acquire 
at least the fundamentals necessary to this end. There seems to be an impression that 
the special knowledge required can be imparted simply by a study of comparative 
machine design and of thermodynamics. The fact is that the origin of most of the 
special training required goes back to theoretical machine design and to the practical 
design problems and exercises connected therewith. Both courses of study therefore 
urgently call for adaptation to the proven need. With this aim in view it is hoped 
that the present work may in a measure serve as a guide and that its many examples 
and problems may furnish welcome material for practice in the drafting-room. 

Augsburg, December, 1902. 


AUTHOR’S PREFACE TO THE SECOND EDITION 


The good reception accorded the first edition of this handbook compelled the 
preparation of a second edition as far back as June, 1903, only a few months after 
the first appearance of the book. On account of business engagements, however, the 
revision for the edition was delayed a very considerable time, although the delay 
had one good feature in that it allowed of the insertion of the latest developments 
in the gas-engine field. The last few years especially have brought forth not a little 
of what is new and of general importance in the construction of gas engines. 

The gas turbine, more than one hundred years after its first appearance, has again 
stepped into the foreground and is again the subject of energetic activity of investi¬ 
gators and inventors, who are apparently blinded by the rapid commercial success of 
the steam turbine. 

Suction gas 'producers, constituting the simplest and cheapest of power installations, 
are in a fair way of completely displacing the older types of gas producers and are 
even making considerable inroads into the field hitherto reserved to the smaller 
illuminating gas and oil engines. This development has incidentally been the cause of 
the taking up of the most important problem yet remaining in the gas-engine field, 
that of the gasification of tar-carrying fuels. This problem is now receiving great 
attention and many fairly successful solutions have already appeared. 

The development and introduction of the double-acting principle in the building of 
4-cycle engines has given the latter type of machine a new lease of life in the struggle 
between it and the 2-cycle engine for supremacy in the field of large engines. The 
eonditions are further improved in favor of the 4-cycle engine by substituting for the 
old cam-operated gear an eccentric valve gear, in part so designed as to operate the 
valves positively throughout. More perfect methods of speed regulation for this type 
of engine have also been devised. 

In this manner one improvement has followed another since the first appearance 
of this bock and this active development of the internal-combustion engine was duly 
taken into account in the preparation -of the second edition. As a consequence, 
Part III, treating of the theoretical design of the various engine parts, and Part IV, 
dealing with modern types of engines, have been considerably enlarged. In order to 
keep the book from becoming unwieldy on account of the necessary insertions, the 
historical discussion of Part I 1 has in this edition been cut down to include only the 


1 This part is entirely omitted from the translation. 


vii 




AUTHOR’S PREFACE TO THE SECOND EDITION 


viii 


distinct and original types of internal-combustion engines and this part has thus been 
shortened in spite of the fact that a brief discussion on gas turbines has been inserted. 
As far as the rest of the main divisions are concerned, no reduction of any importance 
proved possible, in fact, in most of them many, and in some cases extended, insertions 
were urgently called for. 

For the details of the revision the reader is referred to the Table of Contents and 
to the Index; it will be sufficient merely to mention that the number of pages has 
been increased by 80, the number of illustrations in the text by 50, the number of 
plates by 18, and the number of tables by 13. 

Of course neither the inherent character nor the clearly defined aim of the book 
have been changed in any manner. As in the past, its one guiding principle is “ less 
invention, more rational design,” in other words, let us have less guesswork, less rule-of- 
thumb and more of sane design based on tried, principles. And of hardly less import¬ 
ance is the second plea to establish first a sound working basis for rational design before 
■proceeding to a lot of hypothetical thermodynamic experimentation. 

To the first of these demands there are probably few dissenting voices, except 
those of professional inventors. There is, however, no such unanimity of opinion 
regarding the second, which is mainly aimed at our present-day method of technical 
instruction and at scientific investigation. It came as a surprise to note that some 
of the strongest objections to this demand came out of the camp of the practical 
engine builder, while one of the most eminent of our university professors has turned 
into a most energetic champion of the cause opposed to mere theorizing. 

The opinions which the author himself holds concerning one or the other of the 
requirements or demands above laid down, have been clearly set forth in the preface 
to the first edition. The latter expresses a warning against the over-valuation of one¬ 
sided theoretical investigations and merely suggested that a study of the mechanical 
and kinetic laws underlying the construction of the gas engine be in future preferred- 
to the former type of investigation. That, however, is by no means equivalent to a 
condemnation of thermodynamic study in general, much less does it amount to an 
expression of disrespect aimed at the authorities in that field of knowledge, as several 
of the reviews of the first edition were pleased to state. The author desires to leave 
no doubt in any one’s mind that this interpretation does not represent his own ideas 
on the subject. The theory of gas-engine construction is indebted to thermodynamics 
for a great many things and at all times is dependent upon its cooperation,—but 
certainly not to the exclusion of everything else. 

The last words are due also in this case to the colleagues and firms who have 
kindly aided in the preparation of this second edition and who have given of their 
time sometimes at a loss to their own material advantages. To all of them the author 
hereby expresses his thanks, hoping that he may have their continued cooperation. 

Munchen, June, 1905. 


TRANSLATOR’S PREFACE 


In adding this work to the technical gas engine literature already existing in this 
country, the translator feels it hardly necessary to give any extended reasons for his 
belief that the book will fill a real want. If any justification for its appearance 
should be thought necessary, the reader is referred to the author’s preface to the first 
edition. The statements there made concerning the state of technical gas engine 
literature in* Germany apply with equal force to American conditions. As a matter of 
fact, outside of a considerable fund of practical information concerning the automobile 
engine, there is very little data available in our technical literature treating of the de¬ 
sign and construction of stationary engines, large or small. This fact has been brought 
home to the translator in a number of instances, both in his capacity as teacher and as 
consulting engineer or designer. It is by no means intended to convey the impression 
that American gas-engine industry is behind the times or that rules-of-thumb are 
supreme in its designing rooms. Detailed practical data and information no doubt exist 
here as well as abroad, but they have usually been acquired by hard work in the way of 
experimentation at a considerable expenditure of time and money, and are conseqently 
in most cases considered a valuable asset of the business not to be published broad¬ 
cast. There is no doubt that this procedure is entirely justified from the standpoint 
of the manufacturers, but, as Giildner has pointed out, it leads to a large amount of 

useless investigation, results in the appearance on the market of freak forms which 

cannot maintain themselves for any length of time, and, finally, it serves to retard 
standardization of practice. 

It remains to point out in what particulars the book as here presented differs 
from the scope and contents of the original. In the first place the translation is quite 
free without losing the sense of the original. In the attempt to avoid the stilted 

expressions sometimes found in translations, the text was read by at least two engineers 

little conversant with German, and while it is of course quite impossible to avoid 
them altogether, it is believed that so-called “Germanisms” are at least of infrequent 
occurrence. In order to make the book more acceptable to those engineers and 
designers who have had no opportunity to become familiar with the metric system of 
units, all of the figures in both text and illustrations have been transposed from metric 
to English units, except perhaps in a few isolated instances where the system of units 
made no particular difference. To avoid confusing the text, the original metric 
dimensions and figures have not been repeated, as is sometimes the practice. Again, 
there are a few exceptions to this statement in cases where a juxtaposition of the 
figures in the two systems seemed important. The transposition from millimeters and 

ix 



X 


TRANSLATOR’S PREFACE 


centimeters to inches is not in a round-numbered ratio, and since in many of the 
computations and illustrations it seemed important to retain the exact dimensions, 
inches will in most cases be found expressed to the second decimal place, that is, in 
hundredths. This may at first seem strange to the American engineer or mechanic for 
whom the smallest division is usually -fa of an inch. But as stated, this method of 
transposition was necessary in many of the problems and cuts, and for the sake of 
consistency it was followed throughout the book. It is not believed that the value of 
either the problem or the drawing is impaired thereby. 

Perhaps the most radical difference from the original consists in the omission of 
the entire first part treating of the history of the gas engine. Thus Part II of the 
German work has become Part I of the American edition, etc. Although Guldner’s 
treatment of the historical development of the gas engine is quite original and valuable 
on account of the citation of many early test results, it was felt that the same subject 
had been so well treated by English writers like Clerk, Donkin, and Robinson, that 
this particular division of the German edition could be omitted without great detriment 
to the rest of the book. A second reason which contributed to this decision was the 
fact that in order to make the book more acceptable to the American reader in general 
it became necessary to add a somewhat extended discussion of the important machines 
on the American market. This insert largely replaces the pages lost by the omission 
of Part I of the original and, without this omission, the book would undoubtedly have 
become somewhat unwieldy. 

The translator is fully aware that the subject matter of the insert on American 
engines is not in any respect up to the standard set by Guldner in his discussion of 
German machines. In simple justice to himself he is compelled to state that, with a 
few exceptions, this is due to the extreme reticence on the part of many manufacturers, 
as already mentioned, to give out any information concerning their engines, and is in 
no sense due to any neglect on his part to get the best information available. 

In conclusion, the translator wishes to express his sincere thanks to those firms 
who have met to the fullest extent his requests for information and to those of his 
colleagues who have given their time without stint to the revision of manuscript and 
text. Special mention is due to Dr. F. E. Junge of Berlin, who first suggested the 
desirability of undertaking the work and to whose help is due a large share of the 
accomplishment of the same; also to Mr. C. F. Hirshfeld, Professor of Gas Engineering 
in Cornell University and to Mr. A. G. Kessler, Instructor in Gas Engineering, for read¬ 
ing the manuscript and the text, and to Mr. G. W. Lewis, Instructor in Machine De¬ 
sign, for his efficient work in preparing the drawings. 

Ithaca, N. Y., August 16, 1909. 


TABLE OF CONTENTS 


PART I 

THE VARIOUS METHODS OF OPERATING GAS ENGINES AND THE GAS ENGINE 

CYCLES 

A. General Considerations . PA °1 

I. Classification of Engines. 1 

II. Thermodynamic Definitions. 2 

1. Mechanical Equivalent of Heat. 2 

2. Specific Heat. 2 

3. Heating Value and Standard Conditions of Fuel. 5 

4. Thermal Work and Efficiencies. 7 

B. The Various Cycles of Operation. 11 

I. The Constant Volume Cycle.♦.. li 

1. Thermodynamic Examination. 11 

2. Practical Considerations. 13 

II. The Constant Pressure Cycle. 16 

1. Thermodynamic Examination. 16 

2. Practical Considerations. 19 

III. Thermodynamic Aspects of the two Methods of Operation. 20 

C. Critical Examination of the Various Cyclic Events . 28 

I. The 4-cycle Events. 28 

1. The Suction Stroke. 28 

2. The Compression Stroke. 32 

3. The Combustion and Expansion Strokes. 34 

4. The Exhaust Stroke. 39 

5. Heat Interchanges. 40 

II. The 2-cycle Events. 42 

1. The Pump Operations. 42 

2. Scavenging Phenomena. 43 

III. Comparison between 4- and 2-cycle Engines. 50 

1. Theoretical and Actual Thermal Efficiency. . 50 

2. Friction Losses of the Machine. 51 

3. Comparison from Standpoint of Design. 52 

4. Concerning the Question of Economy. 53 

xi 

































TABLE OF CONTENTS 


xii 


PART II 

THE DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION ENGINES 


PAGE 

A. Fundamental Considerations.. 55 

I. Less “Invention,” more “Rational” Design. 55 

II. Horizontal or Vertical Type. 58 

III. With or Without Crosshead. 60 

IV. Single or Double Acting Cylinders. 61 

V. Multi-cylinder Arrangements. 63 

VI. Complete Expansion and Compounding. 66 

VII. Ratio of Stroke to Diameter and Speed of Rotation. 67 

VIII. The Standard Indicator Diagram and Allowable Stresses in Materials. 70 

B. Determination of Principal Dimensions. 72 

I. Cylinder Diameter D and Stroke S on the Basis of Thermodynamic Laws. 72 

II. Dimensions D and S according to Air required for Combustion. 73 

1. General Fundamental Equations. 73 

2. Special Equations for the Principal Fuels. 77 

III. Dimensions D and *S according to Practical Capacity Coefficients. 81 

1. Indicated Mean Effective Pressure. 81 

2. Useful Mean Effective Pressure. 83 

3. Specific Capacity. 83 

4. Specific Piston Displacement. 83 

5. Capacity Constant K . 83 

C. General Engine Parts. 85 

I. Beds and Frames. 85 

1. Frames for Vertical Engines, Crank-shaft above Cylinder. 86 

2. Standard Frames for Vertical Engines. 90 

3. Engine Beds. 102 

4. Main Bearings. 113 

II. Cylinders and Jackets. 120 

1. Cylinders. 120 

2. Jacket Wall. 128 

III. Cylinder Covers (Heads) and Stuffing Boxes. 131 

IV. Pistons, Piston Rings and Piston Rods. 140 

1. Pistons for Single-acting Engines. 140 

Pistons for Double-acting Engines. 140 

2. Piston Rings. 147 

Piston Rods. 154 

V. Crank-shafts. 155 

1. General Kinematic and Stress Relations. 155 

2. Friction Losses in Journals. 160 

(o) The Crank Pin, with Reference to Considerations of Strength. 161 

( b ) Crank-shaft and Main Journals. 162 

3. Strength Computations. 163 

(а) For Dead Center Position. 164 

(б) For Position of Maximum Turning Moment. 170 

Multi-throw Crank-shafts .. 181 

VI. Connecting Rods. 188 

VII. Valves. 194 

VIII. Valve Gearing. 204 

1. Driving Gears. 211 

2. Lay Shaft. 212 

3. Cams. 213 

4. Valve Levers. 214 

5. Adjustment of the Valve Gear. 214 






















































TABLE OF CONTENTS 


xiii 

J»AGB 

IX. Fly Wheels. 215 

1. Determination of Fly-wheel Weight. 215 

2. Determination of Dimensions. 232 

X. Governors. 240 

1. Methods of Governing. 240 

2. Construction of Governors. 242 

XI. Ignition Apparatus. 250 

1. Hot Tube Igniters. 250 

2. Electric Igniters. 252 

XII. Pedestals and Foundations. 257 

D. Special Parts For Gas and Oil Engines. 261 

I. Gas Engines. 261 

II. Oil Engines. 261 

1. Carburetors. 261 

2. Vaporizers and Atomizers. 263 

3. Liquid Fuel Pumps. 266 

E. Auxiliaries. 268 

I. Power Gas Installations. 268 

1. Power Gas Producers (Generators). 269 

(а) Pressure Gas Installations. 272 

(б) Suction Gas Plants. 273 

(c) Gas Producer Installations for Fuels carrying Tar. 282 

2. Gas Washers and Purifiers. 288 

II. Starting Apparatus. 290 

1. Hand Cranks. 291 

2. Mechanical Starting Apparatus. 291 

3. Starting by Fuel Mixture. 292 

4. Starting by Compressed Air. 292 

5. Starting by Electricity. 295 

III. Mufflers. 296 

1. Inlet Mufflers. 296 

2. Exhaust Mufflers. 297 

IV. Cooling Arrangements. 299 

1. Cooling Tanks. 300 

2. Surface Radiators. 301 

3. Cooling Towers. 301 

4. Spray Nozzles and Cooling Ponds. 301 

V. Piping. 302 

1. Air Pipe Line. 302 

2. Gas Pipe Line. 302 

3. Exhaust Pipe Line. 303 

4. Cooling Water Lines. 303 

VI. Gas Meters, Gas Bags, and Pressure Regulators. 304 

1. Gas Meter. 304 

2. Gas Pressure Regulators. 305 

3. Gas Bags. 305 

VII. General Machine Parts .. . 306 

1. Screws and Studs. 306 

2. Keys. 308 

3. Gas Pipe, Pipe Fittings and Flanges. 309 

4. Helical Springs. 312 






















































XIV 


TABLE OF CONTENTS 


PART III 

CONSTRUCTION, ERECTION AND TESTS OF MODERN INTERNAL-COMBUSTION 

ENGINES 

PAGE 

A. Stationary Engines. 313 

I. Capital Cost and Cost of Erection. 313 

1. Fuel Costs. 314 

2. Cost of Attendance. 315 

3. Cost of Lubrication and Cleaning Material. 316 

4. Interest and Depreciation, Computed on First Cost of Engine. 316 

5. Interest and Depreciation, Computed on First Cost of Building. 316 

6 . Maintenance and Repair of Engines and Auxiliaries. 318 

7. Maintenance and Repair on Building . 318 

II. Types of German Engines. 324 

1 . Gasmotoren Fabrik Deutz. 324 

2. Gebr. Korting. 334 

3. Vereinigte Maschinerifabrik Augsburg und Maschinenbau-Ges. Niirnberg A.-G., 

Werk N urnberg. 344 

4. Deutsche Kraftgas-Gesellschaft (Oechelhaiiser). 345 

5. Louis Soest & Co. 353 

6 . Giildner-Motoren-Gesellschaft. 354 

7. Motoren Fabrik Oberursel (Gnom). 361 

8 . Verein. Maschinenfabrik Augsburg und Maschinenbau Ges. Niirnberg A.-G , Werk 

Augsburg (Diesel). 362 

III. Other Types of Foreign Gas Engines. 367 

1 . Langen & Wolf, Vienna. 367 

2. Societe anon. John Cockerill, Seraing. 367 

3. Compagnie “Duplex,” Paris. 374 

4. Societe anon, d’exploitation des brevets Letombe in Lille. 376 

5. Premier Gas Engine Co., Nottingham. 377 

6 . Ganz & Co., Budapest. 379 

IV. American Gas Engines. 385 

1 . Westinghouse Machine Co. 385 

2. De La Vergne Machine Co. 408 

(1) Korting 2-cycle Engine. 410 

( 2 ) Korting 4-cycle Engine. 410 

(3) Hornsby-Akroyd Oil Engine. 416 

(4) De La Vergne 2-cycle Oil Engine. 421 

3. American Diesel Engine Co. 429 

4. Snow Steam Pump Works. 445 

5. Olds Gas Power Co. 451 

Olds Gas Engines. 451 

Olds Gasoline Engines. 457 

Olds Suction Gas Producers. 466 

6 . Riverside Engine Co. 469 

7. Mesta Machine Co. 473 

8 . Allis-Chalmers Co. 475 

9. William Tod Co.... 481 

10. A. H. Alberger Co. 486 

11 . Struthers-Wells Co. 490 

B. Portable and Self-Propelled Engines. 493 

I. Portable Engines. 493 

II. Motor Plows. 493 

III. Motor Vehicles.. 50 i 

IV. Motor Locomotives and Gas Railways. 506 



















































TABLE OF CONTENTS 


xv 


PART IV 

THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 

A TU PAGB 

A. Fuels. 511 

I. Fuel Gases. 514 

1. Illuminating Gas. 514 

2. Oil Gas. 516 

3. Power or Producer Gas. 517 

4. Blast Furnace Gas.. 52 1 

5. Coke Oven Gas. 524 

6 . Brown Coal or Lignite Gas. 525 

7. Acetylene. 526 

8 . Natural Gas. 526 

II. Liquid Fuels. 526 

1 . Crude Oil and its Distillates. 526 

2 . Alcohol. 530 

III. Fuel Mixtures. 533 

B. Combustion in the Gas Engine. 000 

I. Theroretical Data. 539 

II. Older Views concerning Combustion in Gas Engines. 541 

1 . Stratification of Charge and Retarded Combustion. 542 

2. For and against Dissociation. 544 

3. Uniform Mixture and Rapid Combustion: For and Against the Use of the Explo¬ 

sion Port. 545 

III. Current Opinions regarding the Process of Combustion. 549 


APPENDIX 

A. Theory . 563 

I. Synopsis of Thermodynamics. 563 

1 . Heat and Temperature and their Units. 563 

2 . Pressure, Density, and Specific Weight or Gravity. 564 

3. Expansion, Absolute Temperature, and Specific Heat. 566 

4. Relations between Volume, Pressure, and Temperature; Equations of State or 

Condition. 568 

5. Examination and Construction of Pressure-Volume Curves. 571 

5. Relations between Heat and Work, and the Mechanical Equivalent of Heat. 574 

7. Cycles. 576 

8 . “Heat Weight” and Entropy. 579 

II. Fundamental Principles of Thermochemistry. 580 

1. Atoms and Molecules and their Weights. 580 

2. Elements and their Combinations and Symbols. 581 

3. Combining Weights and Volumes, Atomic Heat, Diffusion and Dissociation. 583 

4. Combustion. 585 

5. Distillation and Gasification. 592 

6 . Thermo-chemical Diagrams illustrating various Gasification Phenomena. 598 

B. Some Details from Practice . 610 

I. Directions for Operation, Attendance, etc. 610 

1 . Instruction Book for Olds Kerosene Engines, Type AK. 611 

2. Directions for Starting Hornsby-Akroyd Oil Engines, De La Vergne Machine Co . . 628 

3. Operating Westinghouse Vertical Gas Engines—Instructions to Engineer. 629 

4. Instructions for the Operation of Pressure-Gas Producers manufactured by the 

Gasmotoren Fabrik Deutz. 630 















































XVI 


TABLE OF CONTENTS 


PAGE 

5. Instructions for the Operation of Pressure-Gas Producers made by Korting 

Bros., Hannover. 633 

6. Instructions for the Operation of Suction-Gas Producers made by Gasmotoren- 

Fabrik Deutz. 634 

7. Instructions for the prevention of Accidents in the Operation of Suction-Gas Plants. . 636 

II. Specifications, etc. 637 

1. General Specifications for the Purchasing of Machinery. 637 

III. Regulations concerning the Installation and Use of Internal-combustion Engines. 637 

1. Regulations for the Installation and Operation of Suction-Gas Power Plants. 639 

2. Regulations for the Installation and Use of Gasoline Engines for Agricultural 

Purposes. 640 

I. Stationary Engines. 640 

II. Portable Engines. 641 

3. Special Regulations covering Use of Portable Kerosene Engines in Agricultural 

Industries. 641 

4. Special Regulations covering Use of Alcohol Engines in Agricultural Industries, 

as well as for General Industrial Purposes. 641 

IV. Regulations concerning the Testing of Gas Engines and Gas Producers. 642 

1. Rules for Conducting Tests of Gas and Oil Engines. 642 

2. Rules for Testing Gas Producers and Gas Engines. 656 
















LIST OF TABLES 


NO. _ _ PAGE 

1 . Variation of Specific Heats according to Mallard and Le Chatelier. 3 

2. Effect of Variation of Specific Heat. 4 

3. Thermal Efficiencies Vt of the Constant-volume Cycle for Various Values of e and x . 13 

4. Thermal Efficiencies (Vt) of the Constant-pressure Cycle for P 2 = 467 lbs. and Varying Values 

of p, e, and x ... 19 

5. Comparative Values of Vt and V for Various Methods of Operation. 21 

6 . Compression Temperatures for Various Compression Ratios and Various Initial Pressures... 32 

7. Table of Constants for Gas Engine Fuels. 76 

8 . Capacity Constants for 4-cycle Engines. 81 

9. Temperatures of the Inner Cylinder Wall of Korting Engines. 121 

10. Principal Dimensions, in Inches, of Schwabe’s Gas Engine Stuffing-boxes. 139 

11. Proportions of Snap Rings. 149 

12. Values of p and k for Various Values of —, —, and Kb, for Snap Rings. 150 

s s 

13. Weights of Reciprocating Parts per Square Inch of Piston for Various Types of Engines. 157 

14. Piston and Crank Travel, with the Trigonometric Functions Involved. 160 

15. Table of Constants for Computing o r from the ratio of —. 168 

<Jj 

16. Principal Stresses in Crank-shaft of a 100 H.P. Guldner Engine. 180 

17. Pressures, Speed, and Friction Work in a 100 H.P. Guldner Engine Crank-shaft. 181 

18. Variation of Piston Velocity (0) for ^ = 1:5. 201 

19. Dimensions of Teeth of Screw Gears. 211 

20. Average Dimensions of Valve Gear Parts. 212 

21. Practical Values of p=— . 222 

Vi 

22. Regulation Constants for Various Cylinder Combinations. 230 

23. Required Fly-wheel Rim Weight and Moment GDI for Various Cylinder Combinations.231 

24. Capacity and Dimensions of Beyer’s Spring Governors. 249 

25. Capacity and Dimensions of Hartung’s Spring Governors. 250 

26. Dimensions of Bosch Magnetos. 256 

27. Power Consumption of Starting Motors. 296 

28. Dimensions of Pipe Lines for Commercial Sizes of Engines. 304 

29. Sizes of Gas Meters. 305 

30. Approximate Dimensions of Rubber Bag Pressure Regulators. 305 

31. U. S. or Seller’s System of Screw Threads. 306 

32. Allowable Loads on Screws. 307 

33. Dimensions of Keys. 308 

34. Dimensions of Wrought-iron Welded Pipe. 309 

35. Dimensions of Brass Unions for High-pressure Copper Pipe. 309 

36. Dimensions and Weights of Standard Pipe Fittings. 310 

37. Dimensions of Standard Pipe Flanges. 311 

38. Dimensions of Oval Flanges. 311 

xvii 







































xviii 


LIST OF TABLES 


39. Safe Loads and Deflections of Helical Steel Springs . .. Insert following page 

40. Capital and Operating Costs of Illuminating and Suction Gas Producer Installations. 

41-115 inclusive, Results of Operation and of Tests, Weights, and Dimensions of Modern Types 

of Internal-combustion Engines. 324 

116, 117. Composition and Heating Values of American Coals. 

118. Weight, Heat Density, and Heat Cost of Various Fuels. 

119. Analyses of Illuminating Gases. 

120. Constants for Illuminating Gas. 

121 . Constants for Oil Gas. 

122. Producer Gas Analyses. 

123. Constants for Producer Gas from Coke. 

124. Constants for Producer Gas from Anthracite. 

125. Analyses of Blast-furnace Gases. 

126. Results obtained from Blast-furnace Gas Cleaning Apparatus. 

127. Constants for Blast-furnace Gas. 

128. Constants for Coke Oven Gas. 

129. Constants for Brown Coal or Lignite Gas. 

130. Constants for Natural Gas. 

131. Analyses of Crude Oils. 

132. Classification of Crude Oil Distillates. 

133. Distillation Tests on Mineral Oils. 

134. Quantity of Vapor obtained from Mineral Oils. 

135. Weight-Per Cent Determinations on Alcohol at Various Temperatures Reduced to Standard 

Temperatures. 

136. Specific Volume of Air Under Conditions of Average Humidity. 

137. Explosive Ranges of Pure Air Fuel Mixtures. 

138. Explosive Ranges of Contaminated Fuel Gas-Air Mixtures. 

139. Weight and Heating Value of Fuel Gas-Air Mixtures. 

140. Theoretical Thermal Data for Average Illuminating Gas. 

141. Theoretical Thermal Data for Burned Gases. 

142. Heat Losses through Incomplete Combustion. 

143. Specific Heats and Constant R for Perfect Gases. 

144. Average Values of c p , Cv, and x for Air and Products of Combustion. 

145. Values of Angles a and /? used in the Construction of Curves of the Form PV n = Constant. 

146. Chemical Constants for Perfect Gases. 

147. Combustion Data for the Principal Gases. 

148. Heating Value of the Various Hydrocarbons. 

149. Fundamental Chemical Relations Concerning the Making of Producer Gas. 

150. 151. Forms for Reporting Data and Results of Tests on Gas or Oil Engines.. 


PAGE 

312 

319 

-510 

512 

513 
515 

515 

516 

517 
520 
520 

522 

523 

524 

525 

525 

526 

527 

528 

529 
529 

531 

534 

534 

535 

538 

539 

540 
551 
567 
571 
573 
581 
588 
593 
597 
652 






































LIST OF FOLDING PLATES 


FACING 

PAGE 

Three Piston Pressure and Tangential Effort Diagrams for Single Cylinder Engines. 216 

Two Partial Tangential Effort Diagrams for Multicylinder Engines. 228 

Required Rim Weight of Flywheels in Pounds per I.H.P. 228 

Erecting Plan for Deutz Illuminating Gas Engine (Mod. E 3). 326 

Assembly Drawing, 600 H.P. Deutz Double-Acting 4-cycle Twin Engine (Vertical Cross-section) . . 326 

Assembly Drawing, 600 H.P. Deutz Double-acting 4-cycle Twin Engine. 326 

Erecting and Piping Plan for a Deutz Double-acting 4-cycle Twin Suction Gas Engine. 336 

Erecting Plan for a Korting 2-cycle Blast Furnace Gas Engine. 336 

Assembly Drawing and Erecting Plan, Korting 2-cycle Blast-furnace Gas Engine Direct-connected 

to Blowing Cylinder. 336 

Assembly Drawing, Korting 2-cycle Twin Engine, Rated Capacity 2000 H.P. 336 

Piping Plan for a Double-acting Twin 2-cycle Engine, Korting, for Blast-furnace Gas. 336 

Two Assembly Drawings, 2000 H.P. Nlirnberg Double-acting Tandem 4-cycle Engine. 350 

Erecting and Piping Plan for One of the Older Types of Nlirnberg 4-cycle Blast-furnace Gas 

Engine. 350 

Two Assembly Drawings, Oechelhauser 2-cycle Blast Furnace Gas Engine . ... 350 

Two Assembly Drawings, 1500 B.H.P. Oechelhauser 2-cycle Engine. 350 

Plan of a 1200 H.P. Blast-furnace Gas Installation, Soest & Co., Diisseldorf. 354 

Assembly Drawing, 100 H.P. Giildner Suction Gas Engine. 354 

Plan of a Giildner Suction Gas Plant, Total Capacity 1500 H.P. 354 

Assembly Drawing and Erecting Plan, Diesel Oil Engine. 354 

Plans of Diesel Oil Engine Installations, Total Capacities 400 and 1600 H.P. Respectively. 366 

Assembly Drawing, 3000 B.H.P. Double-acting 4-cycle Tandem Engine, Soc. John Cockerill, Seraing, 

Belgium. 370 

Erecting and Piping Plan of 3600 H.P. Central Station, Soc. John Cockerill, Seraing, Belgium- 370 

Assembly of Triple Cylinder American Diesel Engine. 430 

Installation of American Diesel Engine, U. G. Imp. Co., 3 —16" X 24". 434 

Three Views of Snow Engine, Type B. 446 

Type A, Snow Engine, 400-5000 H.P. 446 




























PART I 


THE VARIOUS METHODS OF OPERATING GAS ENGINES AND THE 

GAS ENGINE CYCLES 


A. GENERAL CONSIDERATIONS 

I. Classification of Engines 

1. The term Internal-combustion Engine is applied to all prime movers employing 
gaseous or liquid fuel, that is, to all gas and oil engines without regard to the cycle of 
operation. Using the cycle as a basis of classification, we have Constant-volume Engines 
in which the charge burns suddenly and at approximately constant volume (Lenoir and 
Otto), and Constant-pressure Engines in which the charge burns gradually and at approxL 
mately constant pressure (Brayton and Diesel). 1 

2. It is not correct to designate a double-acting or multi-cylinder 4-cycle engine by 
the term “2-cycle,” because the expressions “4-cycle” and “2-cycle” state the number 
of strokes required to complete one cycle on one side of the piston. Combining cylinders 
or making one cylinder double-acting in the case of a 4-cycle engine in no way changes 
the number of strokes to each working cycle. 2 As far as the strokes required per cycle are 
concerned, we therefore distinguish only two classes of engines: 

Single-cylinder, single- or double-acting, or multi-cylinder, single- or double-acting, 
4 -cycle engines; and 

Single-cylinder, single- or double-acting, or multi-cylinder, single- or double-acting, 
2 -cycle engines. 

‘Translator’s Note. In German practice, gas engines are sometimes classified as “explosion” 
and “combustion” engines, the former referring to constant-volume, the latter to constant-pressure 
engines. Guldner points out that the classification is not good, for explosion engines are combustion 
engines. He further makes the point that the designation “explosion” engine given to those employing 
the Otto cycle does not exactly meet the case, taking the ground that explosive combustion is much more 
rapid than that occurring in Otto gas engines in ordinary operation. He proposes the term “Verpuffung” 
instead of explosion, meaning rapid combustion but falling short of the nature of an explosion; but since the 
English language does not contain an equivalent term, the name explosion engine will have to stand as an 
alternative for “engines with combustion at constant volume.” 

Guldner objects to this term as well as to the designation “engines with combustion at constant 
pressure,” because of their length; hence the contraction of these expressions into those used in the 
above paragraph. 

2 Translator’s Note. For the same reason, Guldner objects to the term “Ein-Takt,” meaning 
^‘one-stroke cycle,” which is often applied in Germany to the double-acting 2-cycle engine. This term 
has never to the writer’s knowledge been used in the United States. 



2 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 

3. The following distinctions, regarding cylinder combinations, etc., should also be 
drawn at this time: 

As opposed to single-cylinder engines, those having two or more cylinders are known 
as multi-cylinder engines. The latter are subdivided into tandem engines, when the 
cylinders are arranged in a line one behind the other, and double or twin (two-cylinder, 
three-cylinder, four-cylinder, etc.) engines when the cylinders are arranged side by side. 
For other terms designating multi-cylinder engines, see Part II, p. 63. 

Distinguished from horizontal engines, we have the two forms of vertical engines, 
one with the crank-shaft below, and the other with the shaft above the cylinder. As 
usually constructed, horizontal engines are right-handed and run-over, that is, standing 
on the lay shaft side of the engine, the crank will be found on the right-hand side, and 
the direction of rotation is clockwise as viewed from this side. In vertical engines, these 
distinctions are not so clear. Facing what may be termed the front side of the engine 
(the side on which most of the valve gear and the adjustments are found) the wheel is usually 
at the right and the top of the wheel moves toward the spectator. This arrangement, 
however, is by no means strictly adhered to. 


II. Thermodynamic Definitions 

1. The work equivalent of the heat unit (B.T.U.) is taken at 778 foot-pounds, so 
that the heat equivalent of work or the mechanical heat equivalent is 

^4 = -—= .001285 British Thermal Units. 

778 

Many writers still use Joule’s equivalent (772) instead of 778, but the recent investiga¬ 
tions of Dietrich, among others, admit of but little doubt that the true value of the heat 
equivalent is higher, near 783. 1 It is consequently more nearly correct to employ the 

higher of the two values now in common use. With A = B.T.U., thermal efficiency 

computations gives results .935%, that is, less than 1% smaller than if A = B.T.U. 
were used. 

2. The specific heats Cp and Cv may be considered constant for all pressures and 
temperatures occurring in internal-combustion engine practice. This assumption is at 
least safer than the use of the formulae proposed by Mallard and Le Chatelier, Table 1. 
According to these, for example, for a temperature, range from 32 to 3632° F., c v for 
air increases from .1666 to .2506 (i.e., 1:1.51), while for carbon dioxide the increase is 
even greater, from .1422 to .3091 (1:2.18). 

Now recent unimpeachable investigations have seriously shaken faith in the accuracy 
of the above formulae and in the trustworthiness of the experiments from which they 
were derived. Fliegner 2 especially, from his own recent experiments, tersely concludes 
“that the results of Mallard and Le Chatelier may even be looked upon as proof that 
the specific heat of gases at constant volume does not appreciably change at high tem¬ 
peratures, say up to 3600°.” Zeuner, 3 holding the same view, says: “If we were to accept 
the results of Mallard and Le Chatelier as correct, nearly everything at present accepted 

1 Compare, for example, Zeuner, Thermodynamik, 2 ed., Part I, pp. 13 and 120. The International 
Congress of Electrical Engineers in 1904 fixed upon A =1/777 B.T.U. 

2 Vierteljahrschrift der Naturforschenden Gesellschaft in Zurich, 1899, p. 192. 

* Zeuner, Thermodynamik, 2 ed., Part I, p. 143. 


GENERAL CONSIDERATIONS 


3 


or considered established in connection with h'gh temperature computation and with 
the phenomena of the internal-combustion engine would fall. In establishing their 
empirical formulae, these engineers used only very few tests at high temperatures (up to 
3600° F.) of their own, and the trend of their conclusions is therefore at least open to 
question.’* 

Eugene Meyer, 1 in his “Untersuchungen am Gasmotor,” comes to the general con¬ 
clusion that, while the specific heats are not constant, they do not increase with 
temperature as rapidly as stated by Mallard and Le Chatelier. 

A similar stand is taken by Arnold Langen, 2 who, on the basis of numerous scientific 
tests on the explosion pressures of hydrogen and carbon monoxide mixtures in closed 
vessels, derives the following equations: 

For simple (di-atomic) gases, Cv = 4.8 + .0006< 

For water vapor (H 2 0), c v = 5.9 + .00215* 

For carbon dioxide (C0 2 ), c„=6.7-F.00260f 

(Note that t in this case is temperature in degrees C.) 

The final solution of the question whether or not specific heats vary with temperature , 
is of interest not only to the student of thermodynamics, but also, in a high degree, to the 
designer of gas engines. Upon this solution, for example, depends our ability to compute 

Table 1 

VARIATION OF SPECIFIC HEATS ACCORDING TO MALLARD AND LE CHATELIER 


Air. 

Carbon dioxide 
Water vapor . . 

Nitrogen. 

Oxygen. 


Cp — 


C0 2 

h 2 o 

N 

o 


.2360 + 
.2245 + 
.1877 + 
.1421 + 
f .4228 + 
l .3234 + 
/ .2429 + 
l .2312 + 
f .2125 + 
l .2021 + 


.000023* 

.000023 T 
.000093* 
.000093 T 
. 000202 * 

. 000202 T 
.000024* 

.000024 T 
. 000021 * 

. 000021 T 


.1666 + 
.1551+ 
.1423 + 
.0967 + 
.3117 + 
.2123 + 
.1714 + 
.1587 + 
.1500 + 
.1396 + 


.000023* 

.000023 T 
.000093* 

.000093 T 
. 000202 * 

. 000202 T 
.000024* 

.000024 T 
000021 * 
.00002 IT 


the temperatures occurring in the gas engine-cylinder, to determine the energy of the 
working medium going through the cycle, and finally to give some idea of the relation 
between the energy so determined and the various energy losses. As the case stands 
at present, and as long as the question of the variation of specific heat with tempera¬ 
ture is not definitely settled, we are not even able to state positively whether the losses 
due to the imperfections of our present-day engines are relatively large or small, or 
whether, by improving fuel mixtures and ignition, perfecting combustion in general, we 
will be able to raise the thermal efficiency much or little. 

If specific heat increases with temperature as rapidly as indicated by the results 
of Mallard and Le Chatelier, the thermal efficiency of the gas engines of to-day is 
already so high that very little improvement may be expected from anything we can 
do to perfect the thermal action of the cycle. If, on the other hand, the specific heat 


1 Z. d. V. D. I., 1902, p. 1307. 

2 Z. d. V. D. I., 1903, p. 631. 
















4 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


is constant within the limits of temperature found in gas-engine practice, there still 
remains a wide field for improvement in the direction of gas-engine efficiency. 

The following table shows numerically the difference in the results, assuming one or 
the other of the above conditions, regarding specific heat. The experimental results 
were taken from the above mentioned “Untersuchungen” by Meyer. 


Table 2 


EFFECT OF VARIATION OF SPECIFIC HEAT 




Specific Heat Constant. 

Specific Heat Variable. 



Test No. 57 

Test No. 64. 

Test No. 57. 

Test No. 64. 

Temperature of combustion.' 

5 F. 

4746 

3826 

3439 

2960 

Temperature at end of adiabatic expansion.‘ 

>F. 

2856 

2285 

2593 

2153 

Thermal efficiency of engine without losses (Ideal engine) 

% 

39.75 

40.30 

29.69 

31.78 

Relative total work lost referred to the work delivered by 






Ideal engine . 

% 

33.46 

34.50 

10.95 

16.97 

Ratio between the work actually delivered to the possible 






amount of work that may be done by the working fluid 






between the volume limits given. 

% 

66.54 

65.50 

89.05 

83.03 


The values above given refer to a Deutz gas engine having a cylinder diameter of 
8.67", a stroke of 13", and a ratio of compression = 3.84. In Test No. 57 the ratio of 
air to illuminating gas was 8.9 and the total load equal to 10.36 B.H.P. For Test 
No. 64 the corresponding figures were 12 and 6.13 B.H.P. 

These figures show very clearly the importance of the question of the variability 

of specific heat with temperature in its bearing upon the theory of gas-engine design. 

A definite and final answer to this question is not to be expected for some time to 

come. To establish a working basis, therefore, for what is to follow, it is assumed, in 

accordance with the views of Zeuner, Fliegner, and others, that the specific heats c p 
and c v are constant within the ordinary ranges of temperature. The influence of high 
pressures upon the specific heat c P has of late years been made the subject of scien¬ 
tific investigation, mainly by Prof, von Linde. 1 According to the results obtained by 
the latter investigator, atmospheric air shows the following relation: 


Cp — Cpo 1 “f - 


rp2 


^ (T 3 — 3aP) 2 


.]■ 


in which c po = .237; P = pressure in kilograms per square meter; a = 20570 = constant, 
and T is absolute degrees Centigrade. 

The equation gives the following results: 


For P= 10000 

100000 

400000 

700000 Kg./sq. rn. 

p= 1 

10 

40 

70 atmospheres 

= 14.2 

142 

568 

994 pounds/sq. in. 

when T = 273° C. 1 




= 493° F. } c p = * 2375 

.2419 

.2510 

.2779 

when T = 373° C.1 




= 673° F. ! C P~- 2372 

.2389 

.2448 

.2511 


1 Sitzungsberichte der Koniglich Bayer, Akaderaie der Wissensch., 1897, No. 3. 




















GENERAL CONSIDERATIONS 


5 


According to this, the specific heat c p is throughout smaller at 212° than at 32° 
(Mallard and Le Chatelier’s formulae would have shown the opposite result). The 
difference increases with an increase in the pressure. At any one temperature, however, 
c p increases with the pressure, although in a decreasing ratio. For the two pressure 
extremes, p=14.2 lbs. per sq.in. and p = 994 lbs. per sq.in., the increase of c p at 
32° is about 17%, while at 212° the increase is only 6%. Since in gas-engine practice 
the temperatures far exceed 1800° F., we may conclude that Cp for the fuel mixtures 
increases but very little with pressure. Hence the further assumption, for the time 
being, that the specific heats do not vary with pressure is justified. 

Tables giving values of Cp and Cv for the perfect gases and fuel mixtures will be 
found in the Appendix. 

3. Heating Value and Standard Condition of Fuel. In determining the efficiency 
of a gas engine, only the lower heating value of the fuel, i.e., the heat of combustion, 
after subtraction of the heat of vaporization of the water vapor carried by the gases 
of combustion, should be used. The reason for this is that the exhaust temperature 
always far exceeds the temperature of condensation of steam, the latter is discharged 
in vapor form and the heat of vaporization is never available as far as the engine is 
concerned. 1 The difference between lower and higher heating value (Ho) is of course 
proportional to the amount of hydrogen contained in the fud, or to the amount of 
steam carried by the exhaust gases. In the case of illuminating gas, this difference may 
amount to 10% and is therefore quite considerable. 

If one cubic foot or one pound of fuel of heating value H u is mixed with L cu.ft. 
or lbs. of air, the heating value of the mixture will be 

H g = -^y B.T.U./cu.ft. (or B.T.U./lb.).(1) 

l+L 


It is usual to refer heating value and fuel consumption to a fixed set of standard 
conditions. Thus the fuel gases are usually referred to the mean atmospheric pressure 
of 29.92" Hg., and a temperature of 32° F. 2 Thus, if for any given barometric 
pressure of b inches and a temperature of t or T degrees, the lower heating value of a 
fuel is H u B.T.U., this value, by referring it to standard conditions will increase to 


H'u = H t 


29.92 w 461 +t 29.92TH U 


-X 


493 


4936 


rpiT 

~.061^p B.T.U./cu.ft. 


. • ( 2 ) 


On the other hand, any volume of gas v, measured at 6" Hg. and t°, by reference to 
standard conditions, decreases to 

v b 493u „ b _ 493v6 vb 

V °~l + .00203 (i-32)* 29.92 461 +t 29.92 2Q.92T~ T . K ) 

Fuel guarantees for illuminating-gas engines are usually based upon a heating 
value for the gas of 560 B.T.U. per cu.ft. under standard conditions. Suppose now 
that such an engine had consumed v cu.ft. of gas under 6" Hg. pressure and t° tem¬ 
perature, and that the heating value of the gas was H u B.T.U. per cu.ft. under these 


1 E. Meyer, Z. d. V. D. I., 1899, p. 283, in his thermodynamic investigations reaches the same con¬ 
clusions regarding this much discussed point. 

2 For liquid fuel 59° or 60° F. is usually considered the standard reference temperature, but variations 
from this temperature have but a very slight effect upon the fuel consumption of oil engines. 












6 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


conditions. Then the gas volume reduced to standard conditions ot pressure and tem¬ 
perature and to the standard of 560 B.T.U. per cu.ft., will be 


, r v b H'u 

Vo- 16.5 t 56Q - 


But from eq. (2) H’ u = 

^ce .(4) 

It is clear from this that the opposite effects of b and t upon v or H u cancel each 

JJ 

other, and v'o therefore depends only upon 

Example. r = 14.8 cu.ft., H*— 493 B.T.U. per cu.ft. for 6 = 28.55" Hg. and f=53°.6, or 
T = 514°.6 F. 

514 6 V 4Q3 

Then H' u - .061 - = 540 B.T.U. per cu.ft. 

28.55 


Vo 


= 16.5 


14.8X28.55 

514.6 


= 13.5 cu.ft. 


Reducing to the standard heating value = 560 B.T.U. per cu.ft., the volume will be 


v'o = 13.5—— = 13.02 cu.ft. 

560 

The standard conditions to which volumes are most commonly referred, i.e., 
29.92" Hg. and 32° F., are very badly adapted to gas-engine practice because, on this 
basis, all gas volumes are stated smaller, by from 8 to 15%, than the volumes 
actually used or indicated by the gas meter, which latter indication alone determines 
the fuel cost. Fuel consumption computed on the basis of equations (3) or (4) may 
therefore mislead gas-engine buyers or operators as To the real economy of the engines 
under consideration, which fact may in some cases have unpleasant consequences as 
far as the manufacturer is concerned. 

It is of course not open to question that a fixed basis of comparison, i.e., a set of 
standard conditions, is fully as important to gas-engine manufacture and trade as it is 
for scientific investigation; but in order to arrive at such a basis, only such a pressure 
and temperature should be chosen, that the reduced data represent figures as closely 
as possible corresponding to actual results. This is most nearly the case when the 
standard pressure is taken as 28.92" Hg. (=1 metric atmosphere), and the standard 
temperature that already commonly employed as such in chemistry and physics, i.e., 
59° F. ( = 15° C.). In combination with-these figures, the standard heating value for 
illuminating gas may well be placed at H u = 560 B.T.U./cu.ft. 

With this proposed set of standard conditions, eq. (3) may be written in the 
following simple form: 

520v b 17.977 vb \Svb 

Vo 461 + i 28.92 — 461-H ~ T . (3a) 

The approximation in the last form of the above equation causes an error of less than 
T \ of 1% in the exact result. 











GENERAL CONSIDERATIONS 


7 


Example. For 6 — 29.3" Hg., £ = 71°.6, and i> = 31.7 cu.ft, the value of v 0 is from 
Eq. (3) Usual Standard Conditions: 


_31.7X29.3 _ 

V ‘~ 16 - & ~5323 - 2879 eu ft - 

Difference v— v 0 = 9.5%. 


Eq. (3a) Proposed Standard Conditions: 
„31.7X29.3 


r 0 = 18 


= 31.4 cu.ft. 


532.6 

Difference v — v 0 — .95%. 


From the above it would appear that a pressure of 28.92" Hg. and a temperature of 
59° F. represent the most suitable set of standard conditions for gas-engine practice. 

4. Thermal Work and Efficiencies. In judging heat engines from the standpoint 
of economy, it is usual to compare the amount of heat converted into work with the 
amount furnished to the machine in the fuel. This assumes as an ideal case that 
every thermal unit furnished does an amount of work equivalent to 778 ft.-lbs., or 
that each pound of fuel of heating value H gives a return of 

L 0 = ^- = 778H ft.-lbs. 

A 

The duty or capacity of such an ideal engine, when the fuel consumption is C s pounds 
per hour, would be 

T 778 HCs HC S , 

I - = 33600 X 60 = 2545 horSe - pOWer ' . (5) 


The indicated pressure diagram represents the real external work done. Comparing 
this diagram with the ideal pressure diagram of the cycle concerned furnishes a means 
of arriving at the magnitude and cause of the heat losses that occur. 

Fig. 3, p. 11 , shows in full line the ideal pressure diagram of an explosion engine, 
and in dotted line the real indicator diagram. The latter shows the indicated net work 
Li, done by the amount of heat Qi furnished during one stroke. If the work shown 

by the ideal diagram is = ft.-lbs., there will be lost during every cycle an amount 
of heat equivalent to 

= Lt — Li ft.-lbs. j 

or Q_ = AL_ B.T.U. 

This total loss L_ or Q _ is due to a series of causes, such as imperfect ignition and 
combustion, late ignition, and after-burning, conduction of heat during compression and 
expansion, etc., the relative influence of each of which can be judged only approxi¬ 
mately from the indicator diagram. The ratio 7 ^ is called the card factor 1 of the 

Lt 

pressure-volume diagram. 

The area enclosed by the lines above and below the atmospheric pressure line 
depends for its size upon the suction and exhaust resistances, i.e., upon the design of 
the engine and not upon any internal thermal actions. It is therefore wrong, in 
thermal investigations, to subtract the lost work thus represented by the lower loop 
when determining the indicated thermal work Li, as is often done, since the heat 
necessary to make up this loss must come from the heat in the fuel. Only when it is 
desired to determine the purely frictional loss of the machine, as compared with the 


1 In German “ Volligkeitsgrad,” an expression practically untranslatable in a single word. The 
expression "card factor” is much used in steam-engine practice to represent what is practically the same 
thing. 






8 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


total work done against internal resistance, is this method of computation justified. 
Otherwise the subtraction of the fluid loss (lower loop loss) would make the cyclic 
efficiency (iji) less and the mechanical efficiency ( t]m ) of the machine greater than 
they really are. 

If the mean indicated pressure of the indicator diagram is pi pounds per square 
inch, or Pi pounds per square foot, and if the piston displacement is Vh cu.ft., the 
indicated work of one stroke will be 


L* = 144 piVh = PiV h ft.-lbs.(7) 

The corresponding thermal equivalent is 

Qi = A Li = U i V J-J h = B.T.U. 

778 778 

Qi = .185piF/i = .001284PtF a B.T.U.(8 ) 


Assuming that the capacity of the engine is to be Ni I.H.P. at n r.p.m., the piston 
displacement will have to be 


V h = 

for the 4-cycle engine, and 


33000 xNiX 2 
144/w 


458.06A; 
npi 


66000iV» 

nPi 


cu.ft. 


(9> 


_ 33000 X Ni 229.03 Ni _ 33000 Ni 
144 pm npi nPi CU ' 


( 10 ) 


for the 2-cycle machine. 

If an engine uses C s pounds or cu.ft. of fuel per hour of heating value = H B.T.U. 
per pound or cu.ft., the amount of heat actually expended per combustion stroke is 


CsH 

Q = -gL— B.T.U. in a single-acting 4-cycle, 


( 11 ) 


C H 

and Qi = B.T.U. in a single-acting 2-cycle engine. 

The work equivalents of these amounts of heat are 

r Q 77 8CsH 25.93 CsH ^ 4 , 

Li=—= —^-=-- tt.-lbs. for the 4-cycle, 


A 


30n 


and 


L= 


778C S H 12MCsH 
60n n 


ft.-lbs. for the 2-cycle. 


. . ( 12 > 

. . (13} 

. . (14) 


From the above equations the indicated thermal efficiency, in short, the indicated 
efficiency of the cycle (see p. 11) is 

_ ALi Qt NiX 33000 X60 
1)1 Q Q 778 HCs 

2545N, 

..( 15 > 

Dividing the indicated thermal efficiency rji by the theoretical thermal efficiency ry of the 
cycle, we obtain a ratio which expresses the relation between the real machine and 




















GENERAL CONSIDERATIONS 


9 


another machine which is assumed to have no losses. This ratio is again analogous ta 
the card factor , which may then be expressed by 


Tjg = 


Vi 

V 


or using the expressions above developed, for the work equivalents, 



(16) 


(16a) 


A certain part of the indicated work Ni developed by the machine is lost on the 
way from cylinder to shaft through friction of the machine. If we assume that this 
negative (friction) work is equal to N r expressed in I.H.P., the useful (effective) 
engine horse-power will be 

Ne = Nt — Nr . (17) 


The mechanical efficiency of the machine will then be 

Ni — Nr Ne 

r)m -~W~ =s Wi' ' 


(18) 


Substituting for Ni, N e and N r , the corresponding mean pressures pi = p e + pr pounds 
per square inch, we may also write 


Tjm = 


Pi-Pr 

Pi 


Pe 

Pi 


(18a) 


Finally the 
efficiency 


product of the three efficiencies above developed gives the economic 


Tjw = ■rjofjtfjm — IjiTjm. 


(19) 


The economic efficiency referred to the basis of heat consumption may also be expressed 

by 


Niijm X 33000 X 60 2545 N e 
778 HC S ~ HCs ' 


(19a) 


If instead of using the fuel consumption Cs per hour, the consumption C per horse¬ 
power per hour is introduced, we will have the general equation 


2545 

7)10 HC 


( 20 ) 


As determined by experience, V a varies from .40 to .75 according to size and 
quality of engine. Vm = . 75 to .90. Values for Vt are given in Tables 3 and 4, p. 13, 
and 19. 


Example. Vt V g V m V w 
.44 X.72 X.90 =.285 


Vi = . 317 


Energy loss = 1-.285 = .715 = 71.5% of the heat furnished. 


The relation that the various heat losses bear to one another can only be 
determined approximately on account of the uncertainty prevailing with respect to the 
variation of specific heats and dissociation temperatures. It is evident, however, that 
the two main sources of loss are found in the cooling water and in the exhaust gases. 















10 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 



Fig. 1. 


The inter-relation of the heat converted into work and the heat losses is best made 
clear by means of the entropy diagram in which the ordinates represent absolute 

temperatures ( T ), the abscissae the quotient between 
quantity of heat and absolute temperature (Q + T, see 
under Entropy in Appendix), and areas represent quan¬ 
tities of heat ( Q ). The entropy diagram, Fig. 2, has 
been derived from the indicator diagram, Fig. 1, which 
was obtained from one of the later models of the 8 
H.P. Korting illuminating-gas engine. 1 The broken lines 
of Fig. 2 represent the ideal heat diagram that would 
have been obtained with a machine operating without 
loss, assuming constant specific heat. (Regarding the construction and evaluation of 
such diagrams, see Fig. 23, p. 41). The real diagram, shown by the hatched area, 
Fig. 2, represents only about J of the area of the ideal diagram 
(77* = 33.1%, to be exact). Area Fj=34.9% represents the heat 
loss to the cooling water, area F 2 = 33%, the heat lost in the 
exhaust gases. 

The heat lost due to incomplete combustion is included in 
area V 2 . Since the gas particles not burned escape with the 
exhaust gases, this loss by itself is not represented in the diagram. 

But since, as experience has demonstrated, it is possible in 
very unfavorable cases to lose from 10-20% of the heat by 
incomplete combustion, this source of loss is 
by no means negligible. This is of special 
importance since it furnishes the designer with 
a valuable means of improving our present-day 
engines. 

(See also Part II, Valves, and Part IV, 

Combustion.) 

E. Meyer concludes his “ Untersuchungen 
am Gasmotor,” quoted on p. 4, with the fol¬ 
lowing words: “ It cannot be denied that, in 
many engines in operation to-day, especially 
those employing the leaner gases, large quan¬ 
tities of fuel gas pass through the engine unburned, and that the economy of many 
engines can be improved mainly by counteracting the heat losses due to incomplete 
combustion as far as possible by a careful mixing of the new charge.” He bases his 
statement mainly on some of his experimental data given in a table in Part IV 
Subdivision III, of this book. 



- 3000V 


-2500 


~2000 


-1500 


- lOOO 


-soo 


Fig. 2. 


1 From a lecture by C. Linde on “die Auswertung der Brennstoffe als Energietrager,” given before th“ 
Verein d. I., 1903. Z. d. V. D. I., 1903, p. 1509. 










































THE VARIOUS CYCLES OF OPERATION 


11 


B. THE VARIOUS CYCLES OF OPERATION 


I. The Constant=volume Cycle 

1. The thermodynamic examination of the various working cycles is based on the 
assumption, not quite true in actual practice, that the charge, as the carrier of heat, 
is a perfect, ideal gas, which, in a non-conducting cylinder, passes through a reversible 
cycle of operations. With this assumption, the 
resulting pressure diagram is an area bounded by 
four lines, along one of which heat is furnished, 
along another heat is abstracted, while the other 
two are adiabatics. 

If, in such a cycle, Q\ B.T.U. are furnished 
to the gas, while Q 2 B.T.U. are rejected, so 
that Q\ — Q 2 = Q B.T.U. are transformed into ex¬ 
ternal work, then in general the thermal efficiency 
of the cycle is (see p. 8 and Appendix). 


Yl/ Ql — $2 _ ! Q 2 Q 

1 Qi Qx Qi 


( 21 ) 



Now in the constant-volume cycle, whose 
theoretical pressure diagram is shown in Fig. 3, 
the heat furnished during explosion, 

Q 1 =Gc v (T 3 -T 2 ) B.T.U. .(22) 

the heat in the exhaust gases, 


Q 2 = Gcv{T± — Ti) B.T.U.(23) 

hence Q = Qi—Q 2 = Gc v {T 3 — T 2 — T\-\-T\) B.T.U.(24) 


Eq. (24) represents the external work done. G is the weight of the charge 
consisting of Gr pounds of burned gases remaining in the cylinder from the previous 
combustion, Gb pounds of fuel, and Gi pounds of air. Further, c v is the specific heat 
at constant volume, while T and P represent absolute temperatures and pressures for 
points on the diagram marked with the corresponding subscripts. 

Eq. (24) may be written 

«-o a .[r,(i-^)-r,(i-^)] .(24a) 

T T 

From this, bv use of Poisson’s Law, according to which = and bv substituting 

-t 3 ^4 

the value of Qi, eq. (22) for the left half of the expression in parenthesis in eq. (24) 
we finally derive that 


Q = Q 1 


(24 b) 














12 


METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


Hence eq. (21) becomes 

But from the same law, 
therefore we may also write 

V c 


Vt = 

T±= 

t 2 

Vt= 1 




Ql 


= i_n. 

T 2 


VcV- 1 /Pi\*=l 


FA*- 1 


F 


1 


PAulzI 


( 25 ) 


In eq. (25) ^ is the reciprocal of e, the ratio of compression (which in the case 
of this cycle may be taken equal to the ratio of expansion 8). Hence finally 


% = 1-— t =l-£ l ~ x . 


. ( 25 a) 


The same final result is naturally obtained when the derivation of the value for 
the efficiency is based upon the energy equivalent L instead of upon the quantity of 
heat Q. In this case it will simplify matters if the computation is made with the 

specific volume v=^ oi one pound of the charge. 

The absolute energy represented by the expansion stroke is 

. (26> 

The work of the compression stroke is 

.«• 

Subtracting (26a) from (26), the available energy of the theoretical diagram will be- 

r) .W 

Now the energy equivalent of the amount of heat Q furnished the cycle is 

L ~x ° (JV ~ ■ Pa)= (P x~y ^ . m 

From (27) and (28) then 

(Pa-PaW, 1 \ 

s-1 \ s*" 1 / 


L 1 


( P S -P 2 )Vc 
x — 1 


( 29 )’ 


or, as before, 


1 















THE VARIOUS CYCLES OF OPERATION 


13 


Eqs. ( 25 ) and (25 a) show that the thermal efficiency of the constant-volume engine 
depends first upon the ratio of compression e, increasing as the compression pressure P 2 

increases, and second, that it varies in direct ratio with the value of x = —. It is true 

Cv 

that the influence of x is small as compared with that of a variation in e, but the 
difference may amount to 6 or 8% of the minimum value, as shown by the following 

Example. Assume the following constants for two different illuminating-gas mixtures: 

Mixture. Ratio Air to Gas. £ x 

I 1 to 6 5 1.354, then i?t = l-(£) 354 = .435 

II 1 to 13.5 5 1.383, then jj< = 1-(i)- 383 = .461 

„„ .461 - .435 

Difference =-—-= .06 = 6%. 

.43o 

The weaker mixture therefore shows a thermal efficiency 6% higher than that of 
the rich mixture. An additional advantage is that in practical operation, the weaker 
mixture burning at a lower temperature will lose less heat to the jacket water. It is 
therefore advisable in gas-engine practice to use lean mixtures and to compress these 
to a high degree. 1 

Table 3 

THERMAL EFFICIENCIES Tj t OF THE CONSTANT-VOLUME CYCLE FOR VARIOUS VALUES 

£ AND * 


For £ = 

2.0 

2.5 

3.0 

3.5 

4.0 

4.5 

5.0 

6.0 

7.0 

8.0 

9.0 

10 

For *=1.20. 

0.129 

0.167 

0.197 

0.221 

0.242 

0.260 

0.275 

0.301 

0.322 

0.340 

0.356 

0.369 

For * = 1.25. 

0.159 

0.205 

0.240 

0.269 

0.293 

0.313 

0.331 

0.361 

0.385 

0.405 

0.423 

0.438 

For * = 1.30. 

0.188 

0.241 

0.281 

0.313 

0.340 

0.363 

0.383 

0.416 

0.442 

0.464 

0.483 

0.499 

Eor *=1.35. 

0.216 

0.274 

0.319 

0.355 

0.384 

0.409 

0.431 

0.466 

0.494 

0.517 

0.537 

1 0.553 

For *—1.40. 

0.248 

0.313 

0.363 

0.402 

0.434 

0.460 

0.483 

0.520 

0.550 

0.574 

0.594 

0.611 


2. Practical Considerations. The rule just developed does not of course apply in 
practice without qualification. In the first place, contrary to the assumptions above 
made, the real cycle is not a closed one. Further, the expansion and compression 
lines cannot be adiabatic, because the cycle is carried out in a cylinder the walls of 
which are strongly cooled; and lastly, the charge does not behave like a perfect gas 
because chemical changes occur during the cycle. Nevertheless, its accuracy is attested 
by the fact that in practice the indicated efficiency Tji can be considerably improved by 
increasing the amount of compression. In order to produce this net result, therefore, it 
must be assumed that the contradictory influences at work during the cycle operate in 
such manner that those tending to increase the efficiency balance, or rather over¬ 
balance, those tending to lower it. The purely practical advantages of high compression 
are in themselves of some importance; it increases the temperature of the charge 
before ignition, decreases the distances through which the flame must propagate, 
•decreases the external cooling surfaces, and diminishes the proportional admixture of 
4he burned gases to the new charge. All of these effects tend to improve the process 

1 The advantages of the use of lean mixtures, as regards the effect on thermal efficiency, are especially 
marked if it is assumed that specific heats increase with temperatures. See p. 4. 
























14 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


of combustion and to decrease the loss to the cooling water. In the endeavor to 
utilize to its fullest degree this means of improving the economy of internal-combustion 
engines, the compression pressure has been increased step by step from 3 or 4 atm. 
to 8 or 10 (in the case of illuminating gas) and to 12 or 16 atm. (in the case of lean 
gases). Attempts are made to raise the upper limits even beyond these values, but it 
must be remembered that the upper limit is set by the approach of the compression 

temperature to the ignition tempera¬ 
ture of the charge. It is therefore 
interesting to examine into the degree 
of success these endeavors are likely to 
have. 

Fig. 4 shows the increase of Vt with 
a variation in the compression ratio e 
from 2 to 15, assuming that x is con¬ 
stant at 1.35. It will be noted that the 
increase in Vt is less and less rapid, 
and that beyond £ = 15 it increases but 
very little. But, as has been shown 
on p. 9, the practical criterion of the 
value of a machine is not the value 
of Vt, but that of Vw = ir )g r )tVm. The 

effect of variation in the compression 
upon the factor ij g is indirect and is difficult to express numerically. In any case it 
is small, and the injurious effect, for instance, that a lowering of the compression may 
have upon the process of combustion may be easily counteracted by making the 
charge richer and thereby obtaining more effective ignition, etc. If therefore r) g is 

neglected, we may write, for all degrees of compression 

..(30) 

which states that the economy of the gas engine is a function of its thermal and 
mechanical efficiencies only. But while Vt may be accurately computed in all cases, we 

possess absolutely no numerical basis for the determination of rjm. There is up to 

to-day no formula which gives the frictional resistance likely to be encountered in 
internal-combustion engines with even a fair degree of accuracy. We are therefore 
compelled to depend upon experience, which so far seems to have brought out the 
following facts: 

(а) That the mean frictional resistance p r , based on one square inch of piston area, 
increases with the compression, all other conditions remaining the same; and further 

(б) That this frictional resistance p r is either entirely independent of the mean 

indicated pressure pi or varies with it in the inverse ratio. (Compare for instance the 

experimental data in several tables in Part III.) 

To obtain a fairly approximate idea of the value of p r , a curve has been drawn 
in Fig. 4, showing the relation between pr and e for a number of the larger gas 
engines. An examination of the available data shows that, excluding some very 
irregular figures, p T varies according to a law that may be approximately expressed by 

Pr-M\fa lbs. per sq.in.(31) 

The derivation of an absolutely general exact equation fails on account of the 
varying conditions of design and of operation. The relative positions of the p c and pr 



Fig. 4.—Relation between Compression and Efficiency. 



































THE VARIOUS CYCLES OF OPERATION 


15 


curves in Fig. 4 indicate that pr increases at a much slower rate than p c . Now from 
eq. (30), 7j' w does not depend upon p r but upon 


and since pe = pi — pr, that is, since 



Ve_ 

Vi 


fjm — 


Pi-pr 

Pi ’ ’ 


(32) 


it follows that an increase in the mean frictional resistance of the engine does not 
decrease the mechanical efficiency rj m , provided there is an adequate increase in the' 
specific capacity of the engine, as measured by the factor pu To this condition is 
principally due the fact that our modern engines, working with compression pressures 
from two to four times those employed in the first Otto engines, still show a 
mechanical efficiency fully the equal of that found in these early machines. 

Eq. (30) in conjunction with eq. (25 a) gives 


Pi 



Pi~Pr \ 

Pr r 


(30 a) 


from which it follows that a further increase of compression ceases to be beneficial as 
soon as it is obtained at the expense of pi, i.e., as soon as it is rendered possible only by 
making the mixtures extremely lean. 

** The curve for rj m in Fig. 4 was constructed by computing values for the mechanical 
efficiency from eq. (32), substituting values of pi as found from practice. This curve 
shows that up to £ = 6, i.e., p c = appr. 140 lbs. per sq.in., the decrease of rj m is but 
slight, but after that its value drops rapidly. The reason for this decrease is twofold; 
on the one hand it is due to the increase of all the engine dimensions to resist the 
increasing pressures (see p. 68, Part II); on the other, it is probably largely due to 
the fact that the mixtures will have to depart more and more from the normally best 
mixture as the compression temperatures approach that of pre-ignition of the charge. 
The use of leaner mixtures results in lower values of pi, and this, according to eq. (32), 
means a decrease in rj m . 

The product Vthm is finally shown in the curve marked ffw, Fig. 4. Up to £ = 8, 
for which p c = appr. 210 lbs. per sq.in., the rate of increase of rj'w is quite rapid. The 
maximum value is reached for £=10, p c = appr. 280 lbs. per sq.in. Beyond this, an 
increase of £, instead of showing a further increase, shows a decrease in rj' w . We 
conclude from this that in constant-volume engines the use of compression ratios beyond 
£ = 8 does not result in further material gains in the economic efficiency, and that the 
economic limit of maximum compression will lie between 210 and 280 lbs. per sq.in. 
This conclusion does not apply to constant-pressure engines, because for these the 
value of Vt follows a different law, and because rj m decreases more slowly on account 
of the generally greater value of p t (see p. 9). 

The probable maximum value of the explosion pressure p z may be found from the 
following considerations: from p. 34, we have 


Pz = Tz 
Pc To’ 


and 


T Z =T C + 


Q i 

c v Gi’ 


i.e., the increase in pressure is proportional to the temperature increase, and the latter 
is proportional to Q\. High explosion pressures, desirable from the standpoint of 
high specific engine capacity, can therefore be obtained only together with high 






16 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


temperatures, which means with rich mixtures. But such mixtures cause greater 
heat losses in cooling water and in exhaust, to say nothing of some practical 
difficulties. Engines using rich illuminating-gas mixtures with comparatively low com¬ 


pression pressures show an average pressure ratio in the neighborhood of 


— = 4 —5. 

Pc 


The same mixtures used with high compressions would produce explosion pressures 
exceeding 50 atm. and explosion temperatures far in excess of 38*00° F. To avoid such 
dangerous pressures and temperatures, the means generally used is to decrease the heat 
contents of the charge in engines using high compressions, so that the explosion 
pressure shall be from 2.5 to 3 times the pressure of compression, i.e., from 25 to 
30 atm. Under such conditions the dimensions of the engine parts remain within 
reasonable limits, and the heat is much better utilized, as experience has shown. It 
is quite likely, however, that in time the economic compression limit of from 16-18 
atm., which was referred to above, will be approached, resulting in explosion pressures 
in the neighborhood of 40 atm. = app. 575 lbs. The use of the still higher working 
pressures cannot be justified either on the ground of economy or from the standpoint 
of design. 


II. The Constant=pressure Cycle 


1. Thermodynamic Examination of the Constant-pressure Cycle. In the older 

constant-pressure engines (Brayton, Hargreaves, Briinnler, etc.) the taking in and 

compression of the charge, combustion, expan¬ 
sion, and exhaust were carried on in two sepa¬ 
rate cylinders. By combining in Fig. 5 the 



pump diagram a 12 b with the power diagram 
•— o a b 3, we obtain the net energy diagram 
0 12 3. used in the investigation. 

If it is assumed that G pounds of 
burned gases are expanded to atmospheric pressure po lbs. per sq.in., we have, see 
Fig. 5, 


Fig. 5. 


Heat supplied, Qi = Gc P {T 3 - T 2 ) B.T.U.(33) 

Heat discharged, Q 2 = Gc p (T^ — T l ) B.T.U.(34) 

Hence heat transformed is 


Q = Qi-Q2 = Gc p (T 3 -T 2 -T 0 + T 1 ) B.T.U. 

and the thermal efficiency 

r) @ c p(T 3 —T 2 — Tq + Ti) , T(y — T\ 

Gc p (T 3 -T 2 ) ~ L t 3 -t 2 

Assuming adiabatic compression and expansion,- 

To_T l _ T 0 T 3 


rn rji j OF 

i 3 J- 2 

T, 

2Y 

E?|£ 

1 

t-H 

11^ 

S-? 

or 

1 2 


(35) 


(36) 


Hence 


(36a) 

















THE VARIOUS CYCLES OF OPERATION 


1? 


According to Poisson’s law, however, 


and therefore finally 


Tl 

Ti 


x-l 


= cX-1 


= 1- 


Ti 


Ti«* 


r = l- 


ffX-l 




(366) 


The expression for the thermal efficiency of this cycle is therefore the same as that 
of the constant-volume cycle, eq. (25a) showing that the efficiencies of the two cycles 
are the same when the ratio of compression is the same in both. 

It should be observed, however, that for the same cylinder volume or 
piston displacement in each case, the constant-pressure cycle is less 
efficient than the constant-volume cycle because of the necessarily lower 
compression pressure. If, on the other hand, the same maximum com¬ 
bustion pressure be assumed in each case the balance is in favor of the 
constant-pressure cycle. On account of the resulting high compression 
temperatures in the constant-pressure cycle, the latter can then only 
be carried out by separate introduction of fuel. 

This is the principle of the Diesel engine, to 
which Fig. 6 and the following computations 
mainly refer. The final result of the investiga¬ 
tion is not affected in the least if the ignition 
of the charge, instead of being brought about 
by the heat of compression, were caused by any 
other means. 

In constant-pressure engines already con- 

V 

structed, the compression ratio £ — rf is generally greater than the expansion ratio 

V C 

d = The cut-off ratio p — Xf varies with the load on the engine. This fact affects 
Vz Vc 

the efficiency of the process, as shown below. 

With the notation of Fig. 6, 

Qi=Gc p (T- i — T 2 ), .(37) 



Q 2 = Gc v (Ti-T l ), 


(38) 


and the heat transformed to external work again is 

Q = Gc P {T z -T 2 )-Gc v {T^-T l )^Gc P (^ i -T 2 -^^ . . 


. . (39) 


Now 

Hence 

and 

from which 


T 3 =T 2 y = T 2P , 


and T A = T X &) = T lP *. 


Qi=Gc P T 2 (p-l) .(37a) 

Q 2 =Gc v T l (p x -l), .(38 a) 

.(40) 


1 Qi GcpT 2 {p- 1) 


. Cv 1 

But —=—, 

Cp X 


and jr = - 


1 

-X-l > 


hence finally 


Vt = l- 


,o x — 1 




(40a) 
















18 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


Still clearer is the following derivation of the final equation: 


^i_l 

Gcv(Tt-T,) T x I T ! 

H Gc p (T 3 -T 2 ) xT 2 \ 


or, since for the constant-pressure line we can write 

and 


T 3 _V Z T x _(Vc \*- x 

T Vc , and T ( v J , 


we obtain 


^= 1 - 


_T\ 

xTo 


r ir- 

lJH. 




Zl 

xT 2 


TV 


r/ g> - 1 

L.fc-1 


(41) 


(41a) 


1 V 

which, by substituting 7 jT~ = ^zr[, and -y-=p , reduces to the form of eq. (40a). 

The same equation may be derived by using the ratio of the energy equivalent 
L of the heat Q x supplied to the work Lt represented by the theoretical pressure dia¬ 
gram, Fig. 6, as follows: 

y 

Basing the computation again on the specific volume v = 7T of one pound of the charge, 

tr 

we have 

. (42) 

The total energy of the combustion and expansion stroke is 

L a^P 2 (vz-Vc)+y~(l--^ S j .(43) 


The work of compression is found from 


L c = 


P 2 Vc 
X-l 



The available energy then is 


(44) 


Lt - - 1 = | P 2 (Vz - Vc) 



. . (45) 


from which, by dividing by the value of L from eq. (42), it finally follows again that 



Therefore the thermal efficiency of the constant-pressure engine (of the Diesel type), 
besides depending upon the values of the compression ratio e and the ratio of specific 
heats x, also depends upon the cut-off ratio p. Further, the influence of this factor 
is very marked, as may be seen from Table 4, which has been computed from 

eq. (40a), assuming that p 2 = ~ | = 467 lbs., * = 1.30 and 1.41, and p = from 1.50 to 3.0. 















THE VARIOUS CYCLES OF OPERATION 


19 


Table 4 

THERMAL EFFICIENCIES (Vi) OF THE CONSTANT-PRESSURE CYCLE FOR P 2 =467 LBS. AND 

VARYING VALUES OF p, e, AND x 


p= 

1.50 

1.75 

2.0 

2.25 

2.50 

2.75 

3.0 

£=1.3 

£=16 

1^ = 0.535 

0.522 

0.512 

0.499 

0.488 

0.479 

0.471 

£ = 1.41 

£ = 13 

)j< = 0.616 

0.602 

0.588 

0.576 

0.564 

0.552 

0.540 


The inter-relation of the various factors expressed by eq. (40a), i.e., the rule for the 
value of Vt, finds full corroboration in practice in that the indicated thermal efficiency 
Vi of Diesel engines always shows an appreciable increase for decreasing loads, that is, 
for smaller values of p and greater values of x. 

In constant-pressure engines as constructed, owing to the injection of the fuel, there 
is an increase in the charge weight G during combustion. In the case of liquid fuels, 
however, this increase is not considerable and in any case, on account of the variation 
in amount, it would be very difficult to consider its effect in a formula generally 
applicable. Eqs. (37) to (45) therefore do not apply to oil engines with absolute 
rigidity. The exact treatment of the case would give slightly lower values for Vt, 
because, owing to the increase in G, the heat discharged Q 2 will be somewhat greater. 
In the case of constant-pressure engines using gas, this increase in charge weight, 
however, cannot be neglected in thermal investigations, because, depending upon the 
kind of gas, the fuel might under certain conditions constitute one-half of the total 
charge. It is unnecessary to take this matter up in detail at this time since constant- 
pressure gas engines (of the Diesel type) are to-day of no practical importance. 

2. Practical Considerations. The Diesel constant-pressure engine for liquid fuels 
(the only representative of this type of engine) at rated load usually works with a 
cut-off volume equal to 10%, for which p = 2.5. Assuming that a; = 1.41, the thermal 
efficiency of this cycle then is Vt=. 564, from Table 4. A constant-volume engine 
showing the same maximum explosion pressure ( = approximately 467 lbs.) would have 
to show a compression pressure of from 185 to 210 lbs. per sq.in., that is e = 7. 
Assuming again that * = 1.41. the thermal efficiency of this cycle from eq. (25) or (25a) 
would be Vt = . 55. The difference is therefore only 1.5% in favor of the constant- 
pressure cycle. But this advantage disappears with greater values of p, and with 
,0 = 2.75 the balance is already the other way. Based upon the conditions of equal 
maximum pressures in the cycle, which is the only correct basis from the standpoint of 
both design and practice, the two methods of operation are therefore very nearly 

equal. The constant-pressure cycle offers a fundamental advantage only then when its 
maximum pressures of operation far exceed the limit of maximum pressures, set by 
ignition, for the constant-volume cycle. That, however, is not the case in the present 
Diesel engines, because the maximum pressures of from 475-500 lbs. occurring in these 
engines can be easily reached and safely handled in constant-volume engines. The 

undoubted superiority shown by Diesel engines is therefore not so much due to the 
different cycle of operation used by them, but mainly to some purely practical features, 
among which may be mentioned very perfect vaporization of the fuel oil and the 
instantaneous mixing and combustion of the oil vapor. These advantages, however, are 
peculiar to the oil engine only, and would have no bearing if gas were used as fuel. 

For gas fuel, therefore, the constant-pressure principle possesses theoretically no 












20 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


advantage over constant-volume operation; in practice, from the view-point of design 
and operation, the former even labors under a decided disadvantage as compared with 
the latter. 


III. Thermodynamic Aspects of the Two Methods of Operation 

The appearance of Diesel’s pamphlet and, a little later, of his engine, was 
the direct cause of a number of scientific investigations on the various methods 
of operation which did much to clear the old question concerning the fast operating 
cycle. 

To supplement the author’s own view of the question, as outlined in the previous 
paragraphs, the essentials of the most important of the investigations mentioned are 
given below. 

1. Prof. Schottler, in his thermodynamic examination of the Diesel cycle, comes to 
the following conclusion: 1 “Compared to the constant-volume cycle, the efficiency of 
the Diesel cycle is very inferior, unless very lean mixtures and consequently large 
cylinders are used. The latter cycle, however, possesses the advantage of lower 
temperatures and also obtains the benefit of the good effect, already mentioned, that 
lean mixtures have on efficiency. In fact, to the use of lean mixtures may be traced 
the excellent regulation of Diesel engines.” 

Schottler carried out a very extended investigation on the three methods of 

operation (combustion at constant volume, at constant pressure, and at constant 

temperature). Some of the results of his numerical comparisons are given in Table 5. 
In summing up his final conclusion he says: “From all this it follows that the cycle 
introduced by Diesel certainly has undoubted advantages, but that it is quite possible 
to obtain excellent results also by the use of the other processes of combustion. Just 
how the various advantages and disadvantages will manifest themselves in actual 
practice is hard to say off-hand. It should be distinctly mentioned, however, that the 
excellent showing of the Diesel engine is not due to the attempted close approximation 
to the Carnot cycle, but (since pressure rather than temperature is the determining 
factor) that its practical success is a result of the introduction and use of high 
pressures.” 

In Table 5, V stands for the cylinder volumes required by 1 kg. of gas, the 

meaning of the rest of the notation being the same as before. As all figures are 

relative only, this table has not been transposed to English units. A study of the 
figures given again shows that the methods of operation marked (a) and (6) are of 
nearly equal value, while method (c), that is, combustion at constant temperature, as 
laid down in Diesel s Rational Heat Engine,” and covered by his patent, gives 
thermal efficiencies less by from 15-20% than the results obtained by the other 
two cycles under the same conditions. It is therefore not surprising to note that 
the later Diesel machines, operating with combustion at constant pressure, show a 
greater efficiency than the early machines attempting to use so-called isothermal com¬ 
bustion. 


1 Schottler, Die Gasmaschine, 3rd ed., p. 228. 


THE VARIOUS CYCLES OF OPERATION 


21 


Table 5 

COMPARATIVE VALUES OF i)i AND V FOR VARIOUS METHODS OF OPERATION 


Maximum Pressure 50 atm. 


Maximum Pressure 100 atm. 


(a) Constant-volume cycle. 


l F = 


( 6 ) 

(c) 


Constant-pressure cycle 
Constant-temp, cycle.. . 


I P '- 

\ Vt- 
i V- 
I Vt- 
\ V = 


Kgs. of Air to 1 Kg. 
of Gas. 

Kgs. of Air and 
Burned Gas to 

1 Kg. of Gas. 

Kgs. of Air to 1 Kg. 
of Gas. 

Kgs. of Air and 
Burned Gas to 

1 Kg. of Gas. 

50 

100 

50 

100 

50 

100 

50 

100 

0.64 

0.66 

0.59 

0.63 

0.72 

0.72 

0.68 

0.71 

21 + 39 

41+57 

47 

89 

21 + 35 

41 + 54 

45 

O 

00 

1.74 

1 .38 

1.75 

1.38 

1.61 

1.31 

1.61 

1.13 

0.68 

0.68 

0.63 

0.63 

0.74 

0.74 

0.71 

0.72 

21+37 

41 + 57 

45 

87 

21 + 34 

41 + 54 

44 

00 

0.52 

0.61 

0.43 

0.58 

0.64 

0.69 

0.59 

0.67 

21+44 

41 +60 

45 

87 

21 + 39 

41 + 56 

44 

85 


2. Prof. Donat Banki, in a comprehensive article “Zur Theory tier Warme- 
motoren.” 1 comes to the following final conclusion: 

“From all investigations so far carried out it is without doubt necessary to raise the 
pressure in the cylinder of all types of engines as high as practicable. We therefore obtain a 
correct basis of comparison for the different kinds of engines, if we draw our curves with reference 
to the maximum pressures existing in the cycle, that is, existing at the end of the heat supply.” 

A study of the curves cf total efficiency (i.e., economic efficiency rj W ) brings out the following 
facts, important in their bearing upon the design of heat engines: 

The total efficiency increases in all cases with the amount of heat furnished per cycle. The 
curves for rj W) Fig. 7, at first show a rapid rise as the pressures increase, 2 but beyond a certain 
pressure the change becomes more gradual. For a small quantity of heat supplied (180 B.T.U. 
per lb. of air) the curve shows a slight decrease; for a heat supply greater than that, however, 
t]w is steadily increasing in the pressure range investigated. 

The curves for the constant-volume engines rise more rapidly than those for the constant- 
pressure engines; the former therefore intersect the latter. The point of intersection, however, is 
found further into the field, the greater the heat supply. For a supply of 180 B.T.U. per lb. of 
air, for instance, it is located near 10 atm. (140 lbs.), for a supply of 360 B.T.U. it is found 
near 25 atm. (360 lbs.), for 540 B.T.U. near 60 atm. (1000 lbs.), and for 720 B.T.U. the point of 
intersection is above 100 atm. (1400 lbs.). With due reference to practical pressure limits, 
450-600 lbs., we conclude, therefore, that for small supplies of heat the constant-volump engine, 
and for larger supplies the constant-pressure engine, will show the greater efficiency, always pre¬ 
supposing adiabatic compression. 

The curves for isothermal combustion show far lower efficiencies than those for the other 
two engines using adiabatic compression. Besides this, this cycle has the disadvantage of being 
adapted to the utilization of comparatively small quantities of heat only. Hence a machine 
using this method of operation has in any case little to recommend it in practice, and for that 
reason isothermal combustion will not be further considered. 

In engines using isothermal compression, we obtain acceptable efficiency figures only if, in 
the case of the constant-volume engines, the heat supply is at least 360 B.T.U. per lb. of air, 
and in the case of constant-pressure engines at least 540 B.T.U. At a heat supply of 540 B.T.U. 
the efficiency curves for constant-volume engines with adiabatic compression and for those with 
isothermal compression approach each other very closely; for a heat supply of 720 B.T.U. they 

1 Z. d. V. D. I., 1898, p. 893. 

2 position of these curves of efficiency furnishes the explanation for the fact that the earlier 
constant-pressure engines (Brayton, Simon, etc.), all of which operated with low pressures, show an 
efficiency so much inferior to the Diesel engine, in which pressures up to 600 lbs. are used. It also makes 
clear why hot-air engines, in wTiich, besides low r -w T orking pressures, we also have small quantities of heat 
supplied per unit of air, have such insignificant performances to their credit. 


























22 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


nearly coincide. The efficiency curve for the constant-pressure engine with isothermal compres¬ 
sion shows inferior results even for a heat supply of 720 B.T.U., and it is further evident from 
this curve that it is useless to employ compression pressures in excess of 30 atm. (420 lbs.), since 
the losses not considered in this investigation (leaky piston, valves, etc.) would probably over¬ 
balance the slight theoretical gain shown by the curve. 

Assuming that rich mixtures are used, we may say that, concerning the effect upon the 
efficiency of constant-volume engines, it is immaterial whether the charge is cooled during compression 
or not, as long as, through proper choice of compression space, the explosion pressures in all cases 
reach the allowable maximum. Without affecting the efficiency, therefore, it is possible to cool 
to such an extent as to make the compression curve an isothermal. But in such case it is 


% 







BTU's bbb Lb. or/f/n. 





o.so 














360 3 7 

~u 





















b 




0.30 




/8 

OB 

TU 













a 

— 







a 






V 

^4 




~~c~ 




OJO 





c 




h 













( s' 


N 

■A. 







/ 

/ 

- - 

— — 

__e 

— — 

—, _ 






7* 

0.50 

0.30 

OJO 







- r -1- 

3 TU's bbr L s. or A/r. 






> 

5V 

O A 

3T 

u 







720 BTU 









6 

a^. 





_ 









~d~ 







d 






\ 

a/ 


b 



e 




. 

_ _ 



e 






/V 

* 00 








/ / 
_ t/-.i 










/ / 

// 

JL 































A TM. /O 20 30 40 SO so 70 SO 30 /OO /O 20 30 40 SO SO VO 30 30 too 


Fig. 7. —Influence of the Heat Content of the Mixture upon Economic Efficiency Vw- 
a = Explosion engine with adiabatic compression. i> = Constant pressure engine with adiabatic 
compression. c = Engine with isothermal combustion. d = Explosion engine with isothermal compression. 
e = Constant pressure engine with isothermal compression. 


necessary to make the compression space smaller than for adiabatic compression, in order to 
obtain the same explosion pressure. 

For constant-pressure engines, the case is quite different. In this type of machine, cooling 
during compression has an unfavorable effect upon engine efficiency. 

Now, in practice adiabatic compression is not possible on account of necessary cooling of the 
cylinder walls. Hence the efficiency curve of the constant-pressure engine with adiabatic com¬ 
pression will, approach the curve for the same engine with isothermal compression, that is, the 
former null in practice nearly coincide with the curves for the constant-volume engine, or may 
even fall below the latter. 

The results of our investigations may be concisely stated under the following heads: 

a. The supply of heat should be as large as possible (i.e., use rich mixtures). 

b. The method of introducing the heat to the cycle does not affect the efficiency, provided 
the combustion is somewhere between that at constant volume and that at constant pressure. 
















































































THE VARIOUS CYCLES OF OPERATION 


23 


c. In constant-volume engines, the charge may be cooled at will during compression; in the 
constant-pressuie cycle, on the other hand, the degree of cooling should be confined to the 
absolute minimum required for the safety of the cylinder. 

d. The maximum pressures in the cylinder should approach 450-600 lbs. 

3. Prof. E. Meyer, in a very thorough dissertation entitled “Die Beurteilung der 
Kreisprocesse,” does not, like the two investigators previously quoted, start with 
equal maximum pressures in the cycle, but proceeds on the assumption that the 
compression pressures are the same in all cases. He considers each cycle (after the 
manner of Zeuner and others) as the sum of a number of infinitesimal, perfect, 
elementary cycles, each formed by two adiabatic and two isothermal lines; and on this 
basis develops a treatment of great simplicity and clearness which applies with equal 
uniformity to all types of cycles. 1 


From a study of the cycle of Carnot, together with the fundamental facts established by 
the mechanical theory of heat (first and second law of thermodynamics), it is easy to see that 
where work is performed by a quantity of heat falling from a 
higher to a lower temperature level, the amount of work so de¬ 
veloped depends only upon the quantity of heat concerned and 
upon the difference in temperature between the two levels. For 
the purpose of a critical comparison of the various cycles, how¬ 
ever, only that Carnot cycle is suited for which the quantity 
of heat dQ supplied along the upper isothermal (temperature T x ) 
and the quantity of heat dQ 2 discharged along the lower iso¬ 
thermal (temperature T 2 ) are infinitely small, so that the two 
adiabatics are infinitely close together. Under such conditions, it is 
possible to divide any cycle into an infinite number of elementary 
cycles, by means of a series of adiabatics, Fig. 8. With insignifi¬ 
cant error, we may assume that the various heat elements dQ/, 

dQ/' are each supplied at constant temperature, and that the heat elements dQ/, dQ/' are 
each discharged at constant temperature. Then the equations applying to the Carnot cycle in 
general of course apply to the elementary cycles, and we may write for say the nth cycle, 
using Zeuner’s notation, that 



AdL=dQ/ n ) 


T/ri)- T/n) 
T x (») ’ 


and 


9 (») = - 


T/n) -T/n) 

T/ n) 


From this we at once derive the following important rule for the condition of best efficiency 
in a heat engine. Each heat element supplied to a working fluid operating in a cycle should be 
supplied to the cycle at the highest possible temperature, and each heat element discharged from 
the cycle should be rejected at the lowest possible temperature. In other words, the endeavor 
should be to make the temperature limits as wide as possible in each of the elementary cycles. 
This is the fundamental requirement that must be fulfilled. But it should be noted that, in 
general, this requirement does not mean that the total heat 

Q i =d.Q/ + dQ/' + dQ/"+ . . . 


supplied the cycle should be supplied at a single temperature, i.e., along an isothermal with 
finite expansion, and that the total heat 

Q 2 =dQ/+dQ/'+dQ/"+ . . . 

discharged from the cycle should be rejected at a single temperature, i.e., isothermally. This 
would result in a Carnot cycle with finite expansion. Almost without exception the various 


‘Z. d. V. D. I., 1897, p. 1108. 





24 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


heat elements composing the total heat supplied a cycle are not at our command at a single 
constant temperature, and in the endeavor to widen the temperature limits of each individual 
elementary cycle as far as possible, this very fact would lead one to constantly change the 
temperatures of supply and of discharge. Only in the case that all of the heat units supplied 
to and all of those discharged from a cycle are available at constant 
temperature, i.e., when the “sources of heat” are at constant temper¬ 
ature, a case which in the theory of heat engines is the exception, 
may the Carnot cycle be said to be the ideal. 

We will next consider the question how in a gas engine we may 
obtain the greatest possible efficiency, and what method of combustion 
is best adapted to this end. In this discussion it is assumed that the 
gas-engine cycle is a closed cycle, and that the heat Q x is itself in¬ 
troduced into the cycle from an external source, but it is hardly 
necessary here to prove why any of the assumptions are permissible 
without great error. Under these conditions the best criterion for the 
value of the various cycles is again obtained by dividing each into a 
number of elementary Carnot cycles. Let Fig. 9 represent one of these cycles, composed of 
two adiabatics and two infinitesimal isothermals. With the notation of Fig. 9, we may write for 
a perfect gas that 



Fig. 9. 


ii- ir 

Now during the introduction of the infinitesimal heat element dQ u the values of p x and t>, 
undergo but infinitesimal changes, hence we may consider these values the same at the end of 
the isothermal change without great error. The same reasoning applies to the expression 

rp _ PiG \i 

2 R ' 

Therefore the efficiency of the elementary cycle under discussion is 

T x -T 2 p x v x -p 2 v 2 p 2 v 2 

ri = ——— =-= 1 -, 

T i PiVi p x v x 

and finally, since from the adiabatic law PiV^ — p^, we obtain 1 

Pi k 


The above expression, however, may also be considered as having been derived by the use 
of an elementary cycle in which the heat is supplied at a constant pressure p x and rejected at 
a constant pressure p 2 . Depending upon which of the two methods of division gives the clearer 
insight into the case, the choice may therefore be had between this type of elementary cvcle 
or that of Carnot. 

^To prevent any misconception, however, it should be noted at this point that the expression 

2 is derived by aid of the fundamental laws of thermodynamics and the fact that 

temperature is a measure of heat energy, and is consequently applicable to all substances 

fc-i k -1 

without exception; while the equation r=~-- is derived from the former by use of 

Pi k 

the formula pv=RT and the adiabatic law, and is therefore applicable to gases only. Note 


1 The exponent k here used by Meyer is the same as our x=~. 










THE VARIOUS CYCLES OF OPERATION 


25 


especially that for the same pressure limits p, and p 2 , the efficiency jj depends upon the nature 
of the gas, which fact is expressed in the value of the exponent fc=—. The lowest value which 

Cv 

may be assigned to the temperature T 2 is that of the atmosphere; even if the working fluid 
should reach this temperature under a pressure higher than atmospheric, so that if by further 
expansion a lower temperature might easily be reached in the cylinder, this would not improve 
the efficiency. On the other hand, it is of advantage to have the pressure p 2 as low as possible, 
even much lower than atmospheric pressure, as long as the corresponding temperature is higher 
than that of the surrounding air. This is another point that should be carefully kept in mind 
when using the two kinds of elementary cycles. 

From the foregoing discussion it follows that, in order to obtain the greatest efficiency from 
a cycle, it is necessary to supply each heat element at a pressure p t , and to reject the part 

of each element not utilized at a pressure p 2 , so that the ratio — shall be as large as possible. 

Pz 

From this again we may conclude that, given two fixed pressure limits, that cycle is the most 
efficient in which the heat is all supplied at constant pressure at the upper limit and the 
heat not transformed is all discharged at constant pressure at the lower limit, just as the 
Carnot cycle is best for fixed temperature limits. It is therefore quite illogical to assume that 
either one of these cycles can be the ideal in all instances, and to use either one of them to 
determine the available energy for all cases. 

In the case of gas engines there is one other point that should be specially noted. As 
already pointed out, the lower limits of pressure in heat engines at which heat is given off to 
the cooling water, are not in general set by any physical reasons as long as the temperature 
of the working fluid at these pressures is higher than that of the cooling water. In steam 
engines, for instance, in which the lower temperature is produced by water injection into the 
working fluid, the rejection of heat occurs actually under a considerable vacuum. The case of 
the gas engine (with the exception of the atmospheric gas engine) is quite different. In this 
machine a lowering of the pressure in the cylinder below that of the atmosphere is always 
undesirable because the burned gases must be forced out into the air. The production of the 
lower temperatures in gas engines by means of water injection has not yet been tried, and in 
any case the method would seem to offer considerable difficulties. In practice, therefore, in the 
gas engine we have a fixed lower pressure: p 2 can at no time become appreciably smaller than 
atmospheric pressure. Hence the requirement for best efficiency in the gas engine may be 
restated as follows: Each heat element introduced into the cycle must be furnished at the 
highest pressure possible, and the total heat rejected should be discharged if possible at 
atmospheric pressure. This statement explains the great advantage of compression in the gas 
engine. 

The reason for this advantage may also be made clear in another way. Of the two require¬ 
ments for best efficiency in a gas engine, i.e., heat supply at the highest and heat rejection 
at the lowest possible temperature, the former is fulfilled with very great ease. For even at 
atmospheric pressures we obtain temperatures during combustion as high as 2700° F. and these 
in themselves are certainly high enough to guarantee a fair degree of efficiency. On the other 
hand, there is difficulty in rejecting the heat at temperatures that are low enough. At the 
present state of development of the gas engine, low temperatures of rejection can only be 
obtained by allowing the working fluid, after it has attained its highest temperatures, to expand 
adiabatically as far as possible. But since the lower limit to this expansion is set by the 
pressure of the outside air, nothing remains but to supply the heat to the cycle at the highest 
possible pressures. 

Bearing in mind the knowledge thus gained, it is now easy to answer the question as to 
which type of pressure line promises the best efficiency. To this end we compare the process 
of combustion at constant volume, Fig. 10, with that at constant pressure, Fig. 11, and that at 
constant temperature, Fig. 12, and it is assumed in each case that at the beginning of combus¬ 
tion the mixtures shall have the same temperature T c and the same pressure p c . It is further 
assumed that in each case the expansion is carried to atmospheric pressure. Now divide each 
one of the cycles by a series of adiabatics into an infinite number of elementary cycles (for 
the purpose of clearness only three of these are indicated in each case). Each member of the 
series of elementary cycles receives consecutively the equal heat elements dQ /, dQ", dQ ... dQ^ n \ 
An examination of the figures then discloses the following: 


26 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


In the process of combustion at constant volume, each heat element dQi, dQ / f . . . , intro¬ 
duced into the cycle, is utilized to raise the temperature and especially the pressure for the 
introduction of the next succeeding element. Since, however, in all of the elementary cycles, 
expansion is carried to the same final pressure (atmospheric) this process of combustion tends 
of itself to widen the pressure limits for each succeeding cycle, and thus each elementary cycle 
has a higher efficiency than the one preceding it. 

In the case of combustion at constant pressure, only a part of each heat element dQ /, 
dQ/' ... is utilized to raise the temperature for the next succeeding element. But since the 
pressure at which all of the elements are introduced remains the same, the efficiency also 
remains constant from elementary cycle to elementary cycle, each efficiency being the same as 
that of the first elementary cycle in the cycle using combustion at constant volume. 

The case is even less favorable for combustion at constant temperatures. Here during the 
introduction of each heat element the pressure rapidly decreases, so that the pressure limits 
become narrower from cycle to cycle. The efficiency of the elementary cycles therefore constantly 
decreases as combustion proceeds because the ratio of expansion is growing constantly less. 

It follows directly from what has been said that, starting with equal compression pressures, 
combustion at constant volume shows by far the better efficiencies, while combustion at constant 
temperature shows very inferior results. 



It is also easy to see, with the notation of Figs. 10, 11, and 12, and assuming that 
dQi =dQ"—dQi"=dQ^ n \ that the following equations must hold: 

dQ 2 '*=8Q 2 ' = AQ 2 ’, 

dQ 2 '>dQ 2 ”> dQ 2 "'... ><%(*>> ..., 

8Q 2 ' = 8Q 2 " = 8Q 2 " — 

A Qi <4Q 2 " CJQ/" <... <..., 


and from these it again follows that combustion at constant volume is more efficient than 
that at constant pressure, and this in turn gives better results than isothermal combustion. 

On account of the fact then, as shown above, that, for any given compression pressure p c 
combustion at constant volume gives the best results, our aim should be to make both p c and 
lc as high as possible, so that even the first heat element dQ / shall be introduced at high 
pressure, and then even the first elementary cycle will show a high efficiency. But if the 
compression is constantly increased, we soon reach limits set by conditions of actual practice. 
We must therefore next consider both pressure and temperature limits, and such considerations 
will to a great extent change the views obtained on purely theoretical grounds. Temperature 
limits however, play but a subordinate part. Experience has shown that a gas engine may be 
successfully operated even if the temperatures developed in the cylinder exceed 2900° F. Of 









THE VARIOUS CYCLES OF OPERATION 


27 


course with such temperatures it is necessary to thoroughly cool the cylinder bore, valves, seats, 
etc., and this introduces another direct loss of heat, which, however, is of an entirely different 
nature from the quantity of heat necessarily discharged at the lower temperature limit, since by 
it a certain quantity of heat is even prevented from entering the cycle. But, in spite of this, 
if temperatures occur along the compression line in excess of the ignition temperature of the 
gas mixture (appr. 1000° F.), it becomes necessary to cool if it is desired to develop an 
adequate quantity of heat in the cylinder. Schroter has shown that the heat losses to the 
cooling water in Diesel engines are about the same as these losses in constant-volume engines 
(approximately 40% of the heat of combustion), in spite of the fact that in the former the 
temperatures are kept uniformly lower. But, to repeat what has been said, the upper tempera¬ 
ture limits do not affect the general result in any marked degree, because they are, in general, 
certainly high enough to warrant a much better efficiency than we obtain at the present time 
if it were only possible to depress the lower temperature limit. And since, to succeed in the 
latter respect, high initial pressures are necessary, the pressure limits are of the greatest 
importance. The maximum pressure in the cycle is the factor that determines in the first place 
whether any given design is constructively possible or not, whether the machine is cheap or 
costly to build, whether it is easy or difficult of operation, and whether it is sufficiently reliable 
in operation. The pressures at present occurring in Diesel engines may be considered to give an 
idea based on practical experience, of how high we can at present carry maximum working 
pressures. In this connection there is a further difference between temperature and pressure 
limits How high the temperature limit may be put depends largely upon the time during 
which the maximum temperatures last. The shorter the time, the higher may be the limit for 
temperature effects are a function of time. On the other hand, all of the machine parts must 
be built to be safe under the maximum pressure, whether the duration of this pressure is 
instantaneous only or distributed over a longer period. It may be said to be the fact even 
that pressures lasting only an instant are more severe than the same pressures acting during a 


longer interval. . . 

Now suppose that, instead of assuming equal compression pressures m all cycles, we assume 
equal maximum pressures. Then for combustion at constant volume, the pressure at the end 
of combustion must be equal to this maximum pressure. Under these conditions only the last 
of the heat elements is introduced at the most favorable maximum pressure, while all of the 
other elements are introduced at lower, i.e, less favorable pressures. For combustion at constant 
pressure assuming that the working fluid is adiabatically compressed to the maximum pressure, 
each and every heat element is introduced at this most efficient pressure. But it must also be 
evident that combustion at constant temperature gives results very much inferior to either. 
For at the beginning of the isothermal combustion process we have the maximum pressure, for 
which the machine parts must be designed, from which we obtain no advantage, since even 
during combustion the pressure decreases so rapidly that for every succeeding elementary cycle 
the ratio of expansion is considerably less, resulting in decreased efficiency. The net result, 
therefore, assuming a given pressure limit, is that, providing it is possible to compress up to 
this maximum pressure, combustion at constant pressure is the most efficient. Since under these 
conditions the temperature increases during the combustion period it would seem to follow 
theoretically that during this time the loss to the cooling water should be somewhat greater 
than during isothermal combustion; but according to the experience of Diesel who with isothermal 
combustion obtained hardly any development of work diagram, this fact does not have much 


S It is evidently not correct, in answering the very general question as to the most favorable 
cycle for heat engines, to designate certain definite cycles of operation as undoubtedly the best 
ol even “ ideal,” especially as far as the possibility of practical operation is concerned. For the 
choice in every case should depend upon the special governing conditions. The best method of 
procedure for the designer would therefore seem so be to subdivide the pressure diagram which 
he desires to carry out in his proposed machine, into elementary cycles by the method above 
discussed and then bv an examination to see how each elementary cycle might possibly be 
further improved by raising Pl and T„ or by depressing p 2 and T 2 . Thus for instance the ru e 
that combustion at constant pressure is the most efficient does not apply to Otto s 4-cycle 
■ W 31 KJP in this case the allowable limit of compression does not coincide with the 
Lit Tto latter may easily exceed 20 atm,(300 lbs.), but the former must 
not exceed P 7-8 atm. (100-125 lbs.) if pre-ignition in the gas-air mixtures is to be prevented. 


28 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


In such a case then the desire to introduce each heat element at the highest pressure possible, 
naturally leads to the constant-volume cycle, since in it pressures increase most rapidly. 1 


C. CRITICAL EXAMINATION OF THE VARIOUS CYCLIC EVENTS 

The preceding pages contain an investigation of the various operating cycles as a 
whole, according to the laws of thermodynamics, i.e., with reference to the complete 
operation of the transformation of heat into work. It next becomes necessary to 
examine into the various events composing each cycle in so far as actions and inter¬ 
actions occurring are of interest to the designer and constructor. In the following 
discussion, the general notation used in the previous pages has been retained, the 
various states of the working fluid being indicated by proper indices in the accom¬ 
panying diagrams. 


I. The 4=cycie Events 

1. The Suction Stroke, Fig. 13. At the inner piston position, the exhaust stroke 
has just been completed, and the clearance space is filled with V c cu.ft. of burned 
gases under an absolute temperature of T° F., and an absolute pressure of P r in 
pounds per square foot. The weight of this volume in pounds is 


Gr = y r V c = 


VcPr 

TrRr 


(46) 


if y r represents the specific weight of the gases (in pounds per cubic foot), and R r is 
the gas constant. 

If now the piston begins its outstroke, the clearance gases will re-expand until the 
pressure falls from P r to Pa, only after this pressure is reached, fresh mixture (or in 


1 From the general formula for the elementary cycle, Fig. 9, v We may also derive for gases 


\vj V,k-l 


Jfc- 1 


An examination of this equation brings out the following points: In order to utilize the heat 
supplied to a gas engine to the best advantage every heat element should be introduced at the 
smallest volume possible, and the part not utilized should be rejected at the largest volume possible. 
If, therefore, in a gas engine we have fixed volume limits, the most efficient cycle is composed of two 
adiabatics and two lines of constant volume. In the Otto 4-cycle engine we actually have these fixed 
volume limits. The maximum volume is equal to the total volume of the charge drawn in at atmos- 
pheric pressure and temperature; the minimum volume is determined by the consideration of pre-ignition 
of the charge. Consequently the best efficiency is attained when the cycle receives and rejects the 
respective quantities of heat only at the respective volume limits. In the case of the Diesel engine, we 
ave given for the reception of heat a pressure limit, and for the rejection of heat a volume limit 
(since the expansion can proceed only to the volume at which the charge has atmospheric pressure and 
temperature). In this engine therefore, the heat is rejected at constant volume, in which fact it 
differs m its operation from the constant-pressure cycle above discussed, in which expansion is carried 

866 4 underthe *—p— 




CRITICAL EXAMINATION OF THE VARIOUS CYCLIC EVENTS 


29 


the Diesel only air) enters, mixing with the burned gases and thereby becoming 
heated. Assuming that the external temperature is T g , that the barometer pressure is 
30" Hg. (14.7 lbs. per sq.in .) = Pg, and that the 
real volumetric efficiency of the stroke = rj e , 
the mixture drawn in will weigh 

G g = rrjeV h = - - 5 -- pounds. (47) 

1 0 K(i 

At the end of the suction stroke the weight of 
the total charge will be 


Ga = Gr+G g = pounds 

1 alia 


(48) 



_T_i_ 


in which T a is the absolute temperature of ! j /F 
the total charge in degrees F., P a the corre- [ | f 
sponding pressure in pounds per square foot, 
and V=V c +Vh. 

If we assume equal values for the three gas constants, R r , R g , and R a (for 
burned gases, fresh mixture, and for the total charge) which can be done without 
sensible error, we may write 

n t? _VPa _V C Pr 2\\7T) e Vh 

yJaiia ^ T ' T 9 .(48&) 

and from this the actually drawn-in volume of the mixture, referred to a pressure of 
14.7 lbs. per square inch, and a temperature of T a °, is 

T/ , T/ (VP a VcPr\ Tg 

n =-^=(^7—^)2117 cu - ft . (49) 

Distinction should at this point be made between the real volumetric efficiency rj £ 
and what may be called the apparent volumetric efficiency i) V . Referring to Fig. 13, 
the latter may be expressed bv 

Vo 


*)v = 


Fa 


while the former represents the ratio of the volume Vh! of the charge at a pressure P g 
mid a temperature To, actually drawn into the piston displacement Vh, that is 


r)e = 


TV 
Vh ' 


The volume Vh’ is less than V 0 on account of the fact that the temperature in the 
cylinder is greater than Tg. and hence the real volumetric efficiency rj e is always 
smaller than the apparent volumetric efficiency 7) v obtained by ratio of volumes from a 
weak spring card without temperature correction. 

Assuming an atmospheric temperature of t = 62°.5 F., i.e., 7 7 ff = 523°.5 F., the real 
volumetric efficiency r) c of the suction stroke will then be theoretically 


r, e - 


(VPa VcPr\ Tg 
\ Ta Tr / 2117 
V-Vc 


sPa 

T a ‘ 


PA .2A7_ 
'Tr £ 1 


V 


in which e=y- = ratio of compression as before. 


(51) 


















30 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


The value of therefore depends largely upon P a and P r , both of which may be 
easily determined by means of a weak spring. Temperature T r is a function of the 
exhaust and jacket water temperatures, while T a depends mainly upon £ and the 
jacket temperature. If we assume that the gas constants Rr, R g , and R a are equal, it 
follows from eq. (48a) that 

r VPa m 

a GaRa 2117 TjeVh V c Pr 
T 0 + Tr 

and again putting 7^ = 523°.5 F. this reduces to 


T a = - 


.247 VPa 


T}eVh-\- 


.247 VcPr 
Tr 


(52) 


Note that if in any given case the barometer pressure differs considerably from 
30" Hg. (14.7 lbs. per sq.in.), the value of the factor 2117 in the above derivation 
must be altered correspondingly. 

The derivation of eqs. (49)-(52), however, depends upon assumptions and conditions 
which do not quite hold true or exist in actual practice. Thus, for instance, the 
metallic walls act as regenerators of heat; they absorb a considerable quantity of heat 
during combustion and expansion, and transmit a part of this heat immediately to 
the cooling water, transferring the remainder afterwards to the next incoming charge. 
Even if we assume that the mean wall temperature is equal to the outlet temperature 
of the cooling water, there will still be a temperature difference of from 60-90°, 
between the walls and the incoming charge. In reality, however, the walls are much 
hotter than the cooling water, because even from the start a complete transfer of all 
the heat absorbed by them during every combustion and expansion process is not 
possible. The final temperature of the walls at which equilibrium is maintained 
between the heat absorbed and that discharged must therefore be higher than the 
temperature of the cooling water. In small engines with thin cylinder walls this 
difference may be assumed to be at least from 45-55° F., in large cylinders it may be 
from 60-110° F. and more (see p. 121, under "Cylinders ”). On this account the 
temperature difference between walls and fresh charge is likely to be from 150-180°, 
and it is evident that owing to the consequent flow of heat, the charge must be 
considerably heated up, causing a corresponding increase in T a . This action explains 
the well known fact that with increasing load on an engine, the value of jj e decreases. 

Concerning the temperature T r of the exhaust gases remaining in the clearance 
space, we know nothing definitely. It is certain only that it is not lower, but 
probably higher, than the temperature of the gases in exhaust, and T r probably also 
rises and falls with T e . Now since T e is on the average 1540° F. (and over), it 
follows that the clearance gases must even during the suction stroke transfer a con¬ 
siderable quantity of heat to the cylinder walls which are cooler by from 700-900°. 
This action causes a condensation of the burned gases, which in itself affects the 
volumetric efficiency rj e favorably, since the contraction volume practically adds to 
the stroke volume by the same amount. 

The value of r) e is of importance to the designer, especially in the determination 
of the cylinder dimensions required for any given engine capacity (see p. 72, Part III). 
But since it is not possible to determine r? e mathematically, we are compelled to fall 
back on actual measurements of the real volume of mixture W, drawn in per stroke. 








CRITICAL EXAMINATION OF THE VARIOUS CYCLIC EVENTS 31 

In the cas ® of sma11 S as engines such measurements are comparatively easily made 
with the aid of air and gas meters, and they have been occasionally so made in 

scientific investigations; for large engines the volume measurements usually require 
apparatus so bulky and costly that the attempt is not often made. Engineering 

literature consequently offers but little generally applicable information on this point. 
The table below gives a few average values of from reliable data furnished by 

Slaby, Meyer, and others, taken either directly or determined by computation from the 

figures given. 

The remarks made concerning the values of P r and P„ in eqs. (48)-(5I) apply also 
in practice, that is, both should be kept as close as possible to the atmospheric line 
Especially severe are the effects of an increase in P r , since it increases the weight of 
the exhaust gases remaining (by compressing them from P r to P/), and this in turn 
decreases the volumetric efficiency ij e , owing to the re-expansion of the gases from P r 
or P/ to P a . Through a too early closure of the exhaust valve, P/ may be made 
considerably greater than P r with a conse¬ 
quent serious decrease in engine capacity. 

As previously shown, the apparent 
volumetric efficiency may be found from 
the equation 


What values of rjv may occasionally be 
found is well shown by Fig. 14, which 
represents a weak spring card taken by 
the author from an automobile engine. Without reference to any temperature 
differences, this diagram shows a value of ij v = appr. 60%. 

Practical Values of p r , TV, p a , and jj e . For the latter the temperature of the 
charge has been reduced to 32°. 



p pounds per sq.in. 

p a = 12.8 to 13./ lbs./sq.in., = .88 to .93 for slow-speed engines with mechanically 

operated inlet valve. 

/>a = 12.5 to 13.3 lbs./sq.in., ^ e =.80 to .87 for slow-speed engines with automatic inlet 

valve. 

p a =11.8 to 12.5 Ibs./sq.in., ij e = .78 to .85 for high-speed engines with mechanically 

operated inlet valve. 

p a =11.5 to 12.2 lbs./sq.in., 5j e = .65 to .75 for high-speed engines with automatic inlet 

valve. 

Pa= 8.8 to 11.0 lbs./sq.in., ijt=.5 0 to .65 for very high-speed auto-engines with auto¬ 
matic inlet valve and air-cooling. 

Suction generators and carburetors in unfavorable cases may decrease the above 

values of r) e by from 2-5%. 

p r = 15.7 to 17.0 lbs./sq.in. j Too early closure of the exhaust valve and very long or too 

TV = 1180° to 1350° F. 1 small exhaust pipe may increase these values considerably. 








32 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


2. The Compression Stroke, Fig. 13, p. 29. The returning piston compresses the 
charge up to atmospheric pressure and beyond this to the final compression pressure P e 

at the end of the in-stroke. The compression line may be considered as a curve 

following the general law PV n = const. With this assumption, the pressure at the end 
of compression will then be 

Pc = Pa{~X=Pae n , .(53) 

while the final temperature is 

. (54) 


For any given P c or T c the clearance volume must be 



(55) 


From the well known equation PV = RTG, we may also write 


Pc = 


TcGaRa 

~v7~ 


(53 a) 



(54a) 


and 


Vc = 


T cGaRa 

~P^' 


(55a) 


Table 6 

ABSOLUTE COMPRESSION PRESSURES (p c ) AND TEMPERATURES (T c ) FOR DIFFERENT 

VALUES OF £, n, AND T a . 


Ratio of Compression £ = 

2.0 

2.5 

3.0 

3.5 

4.0 

4.5 

5.0 

6.0 

7.0 

8.0 

9.0 

10.0 


For p 0 = 12.5, 

Vc = 

29.7 

39.2 

49.3 

59.9 

70 7 

81.8 

93.2 

117.5 

142.0 

168.0 

195.0 

222.0 



r 600] 


690 

720 

747 

770 

792 

811 

828 

858 

885 

910 

930 

951 



650 


748 

780 

811 

835 

858 

877 

896 

930 

958 

983 

1010 

1028 

n = 1.25 

when T a = ^ 

700 

■ T c = 

805 

840 

875 

900 

924 

945 

965 

1002 

1032 

1062 

1088 

1110 



750 


863 

900 

937 

956 

990 

1013 

1036 

1073 

1107 

1137 

1163 

1188 



UooJ 


921 

960 

1000 

1027 

1056 

1080 

1105 

1144 

1180 

1212 

1240 

1267 


For = 12.5, 

Vc = 

30.7 

41.2 

52.0 

63.7 

75.7 

88.3 

101.2 

128.2 

156.7 

186.5 

217.5 

250.0 



600 


703 

740 

772 

800 

825 

847 

868 

905 

938 

966 

1000 

1018 


when T a = 

650 


763 

803 

837 

866 

895 

917 

942 

980 

1015 

1043 

1076 

1102 

n = 1.30 

700 

T c = 

821 

864 

900 

933 

964 

987 

1012 

1055 

1093 

1128 

1160 

1187 



750 


880 

925 

965 

1000 

1030 

1058 

1087 

1130 

1170 

1210 

1242 

1272 



1800 J 


937 

987 

1028 

1065 

1100 

1130 

1157 

1206 

1250 

1290 

1322 

1355 


For p a = 12.5, 

Pc = 

31.9 

43.0 

55.1 

67.8 

81.2 

95.2 

109.5 

140.5 

162.4 

206.5 

242.2 

300 



f 600' 


718 

760 

797 

832 

860 

885 

912 

955 

995 

1030 

1063 

1092 



| 650 


778 

825 

864 

900 

932 

963 

987 

1035 

1078 

1115 

1152 

1182 

n = 1.35 

when T a = < 

I 700 

Tc = 

838 

888 

932 

970 

1002 

1035 

1064 

1114 

1157 

1200 

1240 

1275 



750 


897 

952 

997 

1038 

1075 

1105 

1139 

1193 

1242 

1284 

1328 

1364 



1800 J 


958 

1012 

1062 

1108 

1146 

1180 

1214 

1264 

1323 

1372 

1418 

1455 


For p a = 12.5, 

Pc — 

33.1 

45.5 

58.8 

73.1 

88.3 

103.0 

120.6 

156.5 

194.0 

234.0 

277.0 321.0 



600 ] 


735 

783 

826 

864 

897 

927 

956 

1013 

1055 

1095 

1135 

1170 

n-1.41 

when T a = ■ 

650 


795 

847 

893 

936 

972 

1003 

1033 

1097 

1140 

1187 

1228 

1268 

700 

■ T c = 

857 

912 

962 

1008 

1048 

1080 

1113 

1180 

1228 

1278 

1320 

1363 



750 


917 

976 

1030 

1077 

1122 

1157 

1194 

1265 

1316 

1368 

1415 

1462 



L 800 J 


977 

1142 

1098 

1149 

1195 

1232 

1273 

1348 

1402 

1458 

1510 

1558 









































































































































































































































34 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


Table 6 , p. 32, diagram Fig. 15, and Plate I, have been constructed on the 

assumption that the vacuum at the end of the suction stroke is .1 atm. (1.42 lbs.), 

that is, that P a = . 9 atm. ( 12.8 lbs.). If in any actual case the real suction pressure 
should be some other value, as Pa, the readings of the above table and diagrams can 
P ' 

be corrected in the ratio - 5 — 

± a 

Concerning the heat interchanges during the compression stroke, we may note the 
following: There is little doubt that the charge temperature T a at the end of the 

suction stroke differs but little from the mean temperature of the cylinder walls. 

During the first part of the compression stroke, as long as the outer, cool end of the 
cylinder is still in contact with the charge, it is likely that the heat of compression is 
completely taken up by the cooling water, keeping the charge temperature about 
constant. But the further the charge is compressed into the inner hot end of the 
cylinder, the faster will the charge temperature increase, and the smaller will be the 
amount of heat taken up by the cooling water. Even if, at the beginning of the 
stroke, there should be a heat transfer from the walls to the charge, such movement 
can be of but short duration and the heat transferred is of small amount. Under 
ordinary cooling conditions, on the other hand, the heat transfer to the jacket is 
always considerable, and this fact manifests itself in the indicator cards in that, with 
increasing compression, the exponent n grows continually smaller, its mean value being 
considerably under the ratio x of the specific heats. 

Values of n from Practice. Owing to the varying heat interchanges along the 
compression line, as pointed out above, the exponent n is not a constant for the entire 
line. Its mean value varies from 1.30 to 1.38 in ordinary cases, with an average of 
about 1.35. But imperfect cooling or high wall temperatures may raise the value of n 
above the adiabatic value x. Charge losses through leaky pistons and valves give 
apparent values of n which are too small. 

3. The Combustion and Expansion Strokes. Diagram Fig. 16. Ignition should, 
at least in theory, begin early enough so that the entire charge is burned at the inner 
piston position. Depending upon the purity, heating value, tempera¬ 
ture, and compression of the charge, as well as upon location and 
nature of the spark, form of combustion chamber, and other practical 
conditions, the time of ignition and combustion is of varying duration, 
as indicated by the more or less inclined position of the combustion 
line or by varying expansion lines. 

The effect of a variation of the point of ignition upon capacity 
and efficiency may be very different in different machines. It is, for 
instance, now and then found that a given machine gives the best 
economic result when the point of ignition is so early as to make 
the beginning of the explosion clearly distinguishable along the com¬ 
pression line, and every explosion is accompanied by a distinct 
knock. Most engines, however, cannot stand a spark so greatly 
advanced, and operate much more smoothly with an ignition point 
moderately late. It is clear that the diagram area, and with it the 
engine capacity, may be increased up to a certain point by late ignition, but the 
reasons for the fact, also sometimes met with, that late ignition favors better utiliza¬ 
tion of heat, are not so patent. 

No definite practical rule for locating the point of ignition can be given. The 
indicator gives reliable information only when the drum receives a continuous motion, 







CRITICAL EXAMINATION OF THE VARIOUS CYCLIC EVENTS 


35 


PLATE I 


































































































































































































































36 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


or when a “displaced” diagram is taken, because in the ordinary indicator card, in 
which the combustion line is placed at the end of the drum motion, the various 
phenomena are not sufficiently distinct. The best means of locating the most favorable 
point of ignition is probably to make the ignition gear adjustable, and then to locate 
the best point by trial. 

E. Meyer, in his “Untersuchungen am Gasmotor,” has the following to say 
regarding this question: 1 “The earlier the point of ignition, the greater the heat loss 
to the walls. The thermal efficiency reaches its maximum with a mean point of 
ignition, but this point does not need to be very strictly maintained because the 
consumption of heat differs but little within the limits of 15° of crank angle, and only 
shows a decided increase for an early point of ignition (20° crank angle below the 
dead center). As for the rest, late ignition causes a loss of work through spreading 
of the indicator diagram, but at the same time the loss of heat to the walls is 
less.” 

It should be remembered that the above remarks apply to a single definite 
engine. 

Assuming that the combustion is instantaneous, that is, occurs at constant volume, 
a condition which hardly ever obtains in practice, the maximum explosion pressure 
will be 


The explosion temperature is 


or, 


p P c Tz TzGaRr 

p T c - VT . (56) 

U Pc GaR* . (5?) 

T ’ = T ' + St .< 58 ) 


when G a pounds of charge contain Q i heat units. Eq. (58), however, cannot be 
depended upon since the results may be affected by incomplete combustion and other 
causes. 

When the combustion becomes slow through late ignition or other causes, so that 
Vc>V c the above equations change to 


and 


P f z = Pc 


T'z=T c 


T’zVc TzGaRr 


TcV'c V'c . 


P'zV'c P'zV'c 


PcVc GaRr 


If the values of P z or P' z can be taken directly from an indicator diagram, 
eqs. (57) or (57a) will give the real combustion temperatures T z or T' e . If, on the 
other hand, P z is taken from a hypothetical indicator card, the real value of T z , or, 
in the reverse case, of P z can only be found by multiplying by a reduction factor 
whose value expresses the decrease of temperature or pressure owing to heat losses, 
cooling, etc., and which value is not far from that of the card factor i) g of the cycle 
(see p. 7). 


1 Z. d. V. D. Ing., 1902, p. 1307, 












CRITICAL EXAMINATION OF THE VARIOUS CYCLIC EVENTS 


37 


At the moment the exhaust valve opens, i.e., at the beginning of exhaust, the 
burned gases have expanded to the terminal pressure (see diagram, Fig. 13). 


Pe = l\ 


VcV Pz 
V 


(59) 


V ' 


in which represents the real ratio of expansion. At the same time the end 

temperature is 

or, using G a and Rr, we also have 




Pe = 


and 


T e = 


T e Ga Rr 

V ' ’ 

PeV' 


GaRr 


(60) 

(59a) 

(60a) 


V 2 * 


It would appear from equations (59) and (60) that the greater the ratio of 
expansion, the lower will be the terminal pressure and temperature, which would make 
the principle of “complete expansion” appear of 
great benefit. In 4-cycle Otto engines, however, 
the ratio of expansion is in most cases nearly equal 
to the ratio of compression, that is, d ~ s, from which 
we may write eq. (59), 


r 




(61) 


This equation shows that an increase in 
compression also causes a decrease of terminal 
pressure and temperature, which explains the 
fact that in our present day of high-compression 
engines, complete expansion has lost all practical 
importance. 

In the constant-pressure cycle, Fig. 17, the compression pressure P c is about equal 
to the maximum combustion pressure Pz. With a cut-off ratio P—^ an d a ratio of 



Fig. 17. 


V 

expansion ^ = y r we then have 


LZ ~ LC V C ° P ~GaRr’ 


or, assuming complete combustion as before, 


Tz — T c + 


Q i 
CpGa 


(62) 


(62a) 


For the end-point of the expansion we may write, as in the case of the constant- 
volume cycle, 


MS)’-* 


(63) 













38 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


and 

or, 


' \W ““d”- 1 ' 


TeGaRr 

V' ’ 



(64) 

(63a) 


and = .(64a) 

In both the constant-volume and the constant-pressure cycles, the real exhaust 
temperature is less than the terminal temperature T s , above computed, because the 
former is measured outside of the cylinder after complete expansion of the gases to 
atmosphere. In both cycles the expansion of the burned gases in the cylinder takes 
place under continued loss of heat to the cooling water, hence the exponents n of the 
expansion line are considerably greater than the adiabatic ratio x of the specific heats 
of the charge. At the beginning of the expansion, when the cooling surface of the 
cylinder is still small and the losses to the cooling water are more or less balanced 
through after-burning of the charge, the differences between n and x are in most cases 
quite small. It may even happen that the heat supply due to after-burning balances 
or overbalances the entire heat loss, in which case n approaches the value of the 
isothermal exponent. But the greater the cooling surface uncovered by the outgoing 
piston, the greater the heat losses and the greater the corresponding value of n, 
although on account of the accompanying rapid fall in the temperature and the 
consequent smaller loss of heat to the jacket, the rise in the value of n is less rapid 
near the outer end of the stroke. It also often happens that the value of n varies 
quite irregularly, alternately increasing and decreasing, so that it becomes impossible to 
recognize any definite law of variation. An increase in the piston speed is, for obvious 
reasons, nearly always accompanied by a decrease in the mean value of n for the 
entire stroke. 

For the method of determining n from given indicator diagrams, see Appendix. 

Values from Practice. The value of the exponent for the expansion line is 
generally found between 1.35 and 1.50, although diagrams showing Up to n = 1.70 are not 
rare. Loss of charge through leakage increases the real value of the exponent; indicators 
having considerable friction loss have an apparent effect in the opposite direction. 

The explosion temperature T z in our present day high compression engines no 
longer remains much below 3600° F. abs. (see the entropy diagram, Fig. 2, p. 10). 

This may be quickly shown by an approximate computation: For £ = 6, n=1.30, 

p a = 12.8 lbs. per sq.in., and T o = 684 0 F. abs., we get from Table 6, p c =9.24 atm. 

(131.5 lbs.) and !Fc = 1170° F. abs. If we assume that — ~3, a figure easy to reach for 

Pc 

395 

maximum load, we have p z = 395 lbs. and from eq. (57) then T z = 1170-—- = ~3500 3 

lo 1.0 

F. abs. The maximum pressure in the cycle is usually also accompanied by the 
maximum temperature, i.e., the heat supplied by after-burning is less than that used 
by the expansion and cooling of the gases. The opposite condition is at least no 

longer common in our modern engines. Assuming a mean exponent n = 1.5 for the 

expansion line, the temperature at the end of the expansion in the case chosen would 

be T e = —~ 1440° F. abs.; for <5=3, however, T e would have been equal to 
6 

3500 ino „ 0 „ 

^ 5 - = approx. 1980° F. 








CRITICAL EXAMINATION OF THE VARIOUS CYCLIC EVENTS 39 


In the case of the constant-pressure engine, if, in agreement with actual conditions, 
we assume p c = pz= 33 atm. (470 lbs.), £ = 16, 3=5.5, p= 3.0, and further take the 
exponent n as above, the maximum combustion temperature will be, from eq. (62), 

4740 

Tz= 1580X3=4740° F. and the end temperature of expansion, from eq. (64), T e = —^ = 

5.5 

2040° F. abs. The maximum and mean temperatures of the constant-volume cycle 
are therefore considerably smaller than those of Diesel’s constant-pressure engine. 

4. The Exhaust Stroke, Diagram Fig. 13, p. 29. The exhaust valve usually 
commences to open at about 90% of the expansion stroke. The burned gases escape 
and the pressure sinks rapidly from P e to P r . During the establishment of pressure 
equilibrium (exhaust) the velocity of the outflowing gases is very high, in the neighbor¬ 
hood of 2600 to 3000 ft. per second, 1 and is independent of the port area. From the 
kinetic energy of the exhaust gases 

^ GaC 2 

Q0 778X32.2X2’. (65) 


in which G a is the weight of the gases in pounds, we may derive the mean exhaust 
velocity 


C = 224\feft. 

Cr a 


per sec. 


( 66 ) 


The simplest way of determining Q a is to consider it as the third member (remainder) 
in the heat balance, the other two being the heat equivalent of the work and the 
jacket loss respectively. 

The actual exhaust should be completed by the time the piston has reached the 
dead center, so that on the back-travel it is called upon to overcome only the 
resistance Pr due to friction in valves and ports. 

If on the other hand the exhaust is retarded (see 
Fig. 18, from a high-speed engine), the effect is 
not only to increase the negative work but also to 
increase cylinder temperature, which in turn seriously 
decreases the volumetric efficiency and the allow¬ 
able compression pressure. Regarding the harmful 
effect of P r , see p. 31. 

If the lower loop of the diagram is taken with a weak spring, remarkable pressure 
variations are sometimes disclosed. In many cases their cause is not to be sought as 
much in the engine as in the exhaust line. The exhaust gases, discharged very 
suddenly and under very high velocity, acquire under certain conditions certain 
vibratory movements accompanied by corresponding pressure variations. With an 
oxhaust line of suflicient length and no sudden turns, the kinetic energy of the exhaust 
gas column, initially acquired, may be sufficient at times to overcome all frictional 
resistances in the line, so that P r drops to atmospheric pressure, or, owing to over- 
expanding of the gas, it may even drop below it. Crossley Brothers tried to make this 
accidental phenomenon one of regular occurrence and attempted to utilize it as a 
means for scavenging the cylinder. Jul. Sohnlein even went a step further, intending 
to cause the vibrating gas column to draw in the new charge, in 2-cycle engines. 2 

1 See Slaby, Kalorimetrische Untersuchungen, p. 191. 

2 German Patent, No. 83210, 1894. 


about 500 r.p.m. 



Fig. 18. 









40 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


Neither one of these schemes had any practical success, because the occurrence of the 
vacuum could not be controlled with certainty, the exhaust action being subject to the 
influences of too many accidental conditions. With ports sufficiently large and with 
correctly timed valve opening, the occurrence of a vacuum at the beginning of the 
exhaust stroke is common, quite probably because at that time the kinetic energy of 
the gas column is greatest (see diagram Fig. 19, from one of the older types of 



Fig. 19. 



Crossley engine); in some other cases this vacuum persists until near the middle of the 
stroke (Fig. 20, from a Diesel engine); or it may not occur at all (Fig. 21, card from a 
Priestman engine). If the frictional resistances encountered in the exhaust pipe are of 
such a nature as to set the gases into vibration, the vacuum may occur intermittently, 
the line cutting the atmospheric pressure line repeatedly (Fig. 22, diagram from an Otto 
engine). It was shown by E. Meyer in the Zeitschrift d. Y. D. I., 1901, p. 1343, that 
this wave form of the exhaust line is not always due to inertia effects in the indicator. 

Values from Practice. The pressure along the exhaust line is usually from 15 to 
16.5 lbs. abs., sometimes more; the exhaust temperature (outside of the cylinder close 
to the exhaust valve) varies from 950° to 1450° F. abs. As might be expected, this 
temperature may be raised considerably by use of rich mixtures higher (values of pi) 7 



Fig. 21. 



Fig. 22. 


late ignition, strong after-burning, etc. Increased piston speed and less effective cooling 
on the other hand raise this temperature but little. 

For values of the temperature T t at the moment of exhaust opening, see p. 38. 

5. The Heat Interchanges between charge and envelope during the individual strokes 
may best be collectively studied by means of the entropy-temperature diagram (see 
Appendix), an example of which has already been given on page 10 in connection with 
the explanation on heat losses. In Fig. 23 the entropy-temperature diagrams (based on 
1 kg. of charge) for a Deutz gas engine (full heavy line), a Grob & Co’s, kerosene 
engine (broken line), and a Diesel oil engine (full light line) have been superimposed, 
the numbers referring to the pressures in atmospheres at the points of location. 1 
Remembering that movement to the right indicates addition of heat, and that to the left 
shows loss of heat, the diagram gives a clear picture of the heat interchanges during 
the second, third, and fourth strokes. Starting with the two explosion engines, the 


1 Krauss, Z. d. Oesterr. Ingr. & Arch. Verein, 1898, No. 10. The figures in this diagram have not 
been transposed to English units, since they are only used for comparison. 


















CRITICAL EXAMINATION OF THE VARIOUS CYCLIC EVENTS 


41 


line a~6 of the Otto engine shows a moderate loss of heat throughout the entire com¬ 
pression stroke; in the case of the Grob kerosene engine, on the other hand, the loss is 
so strong as to make the line ab' appear almost isothermal. (It should be noted, how¬ 
ever, that the very small pressure ordinates of the first half of the compression can only 
be shown approximately, so that this part of the entropy diagram is always somewhat 
uncertain.) At point b or b' the explosion commences, showing strong addition of heat, 
as a consequence of which the curves turn upward and to the right, rising to c and c' 
respectively. Apparently the maximum pressure coincides with the maximum tempera¬ 
ture in both diagrams. The expansion in the explosion engines takes place along the- 
lines cd or c'd f , with constant further addition of heat, due to after-burning, in the gas 
engine, and in the kerosene engine almost 
adiabatically after two thirds of the ex¬ 
pansion stroke is completed. But since 
some heat must be lost to the cooling 
water, there must also be after-burning 
all along the expansion line in the kero¬ 
sene engine. Finally, the exhaust takes *200 
place along the lines da and d'a, during 
which heat is largely abstracted from the 

/ooo 

cycle. 

In the entropy diagram of the Diesel 
engine, line I—III represents the compres- & 
sion stroke. The marked changes of the 
direction in this curve are perhaps not 
absolutely certain, owing to the reason GOO 
mentioned above, but there is no doubt 
that the general trend of the line to the 
left shows that the compression is accom¬ 
panied by a loss of heat. This loss is 
so marked beyond the point II that the 
temperature of the compressed air re- 
mains practically constant. At point III 
injection and combustion of the fuel 
commences, at once the line turns upwards and to the right, showing continued, 
strong supply of heat to the point V. During the same time the temperature rises 
rapidly from 580° to 1510° C. abs., that is, the increase is in the ratio of about 2.6 to 
1. The temperature increase does not stop with the attainment of maximum pressure 
(point IV), as in the case of the explosion engines, but continues beyond this point to 
the point V, where the pressure has decreased to 24 atm. Between V and VI the 
temperature of the mixture remains practically constant (in this particular case); the 
heat used by the work of expansion and that lost .due to cooling is therefore replaced 
by after-burning along this line. From VI to VII the heat supply from the latter 
cause is no longer quite able to fully replace the loss due to the other two causes, 
hence the temperature gradually drops, the change from VII to VIII being apparently 
purely adiabatic. The continued heat loss to the cooling water is therefore also in this 
case replaced to the end of expansion by after-burning. At the beginning of exhaust 
at point VIII the gases still have a temperature of over 1100° C., which, due to the 
heat loss in the exhaust gases, rapidly drops to about 300° C. at the point of 
beginning. 







r 

zr 


1 

'SOTPLRM 



// 

f\~ 



X 

N 


/ 

fc 

1 8 




s 


v 



e 



to 

e/y 

/ 

42 

i c ' 

* \ -- 

V/ 




v! 

0 

/ 


\ 

% 

% 






/of 

V 

1 / 




/ss 



1 / 

3/d 



/ 







in/ 

/ 





0 


3S~ 


7 


2/ 








_ 



f, 

z>> 

Ko 


/ h 

'EAT, 

Supply' 



1 



m=3.9e 
m btu 


23.—Entropy Diagrams for Constant Volume 
and Constant Pressure Engines. 






























42 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


All three of the diagrams refer to engines brought out between 1895 and 1897. 
The changes which the pressure-volume diagram of the Diesel engine has experienced in 
the meantime would manifest themselves in the entropy diagram in that, owing to the 
increased supply of heat (or the increased temperature), the points IV and V, and 
consequently also VIII, would be moved correspondingly higher up. 

In the writer’s opinion, none of these entropy diagrams can claim perfect accuracy. 
That for the Diesel engine probably deserves the greatest credence because, except for 
the specific heats, complete data for its computation were given in Schroter’s report 
on the Diesel engine of 1897. On the other hand, the lines for the diagrams of the 
explosion engines appear to have been determined on the basis of some assumptions at 
least open to questions. But in general the diagrams serve to give a picture of the 
heat interchanges in the cylinder sufficiently correct in its main features, which is all 
that is desired here. 


II. The Events of the Two=stroke Cycle 

In the usual form of 2-cycle engine, in which the exhaust gases are discharged during 
the time that the piston passes the dead center, the compression, combustion, and 
expansion of the charge take place exactly as in 4-cycle machines, so that as far as 
these events are concerned, we may refer to the previous articles. From this point of 
view, the only events belonging peculiarly to the two-stroke cycle are the charging and 
discharging actions of the mixture or scavenging pumps, which actions may continue 
into the cylinders or receivers. Unfortunately, as far as the development of the 2-cycle 
engine is concerned, but little has been done to systematically investigate and clear the 
most important internal phenomena connected with the charging and discharging 
actions of the two-stroke cycle. Inventors and manufacturers who have developed 
2-cycle engines are very reticent when it comes to discussing their good or bad 
experiences with this type of machine; and as for scientific investigation, the two-stroke 
cycle no longer seems to offer to the laboratories of our technical schools any points 
worth clearing up. The natural result is that to-day, as for the last thirty years, the 
designer is compelled to solve all fundamental as well as constructive problems relating 
to the 2-cycle engine for himself, by the aid of trial, experience, and observation. 

Among the problems open to debate regarding the 2-cycle engine might be 
mentioned the following: 

The pressure of the scavenging agent, and the duration of the scavenging period. 

The size and position of the scavenging ports. 

The point of admission of the scavenging agent, whether at the cylinder head or 
near the piston face. 

The question of intermediate receivers, whether any should be used, and if so, the 
proper size. 

The question of excess of scavenging air, whether much or little, etc. 

The following discussion is very largely based upon the experience gained by the 
author himself. How far the conclusions arrived at may be generalized may be seen 
from the appended discussion by A. Wagener regarding the charging and discharging 
actions in large 2-cycle engines. 

1. The Pump Operations offer nothing new in themselves, since the charging or 
scavenging pumps do not differ materially from ordinary air pumps. The suction 
stroke of the 2-cycle engine differs from that of the 4-cycle only in that it is carried 
out in a separate pump, which latter is in all respects designed for this service only. 


CRITICAL EXAMINATION OF THE VARIOUS CYCLIC EVENTS 


43 


The large clearance space (combustion chamber of the 4-cycle engine) is eliminated, the 
pump cylinder is not heated up, and the valves, since they are subject to small 
pressures only, can be made of proper port area and of light weight. All these things 
eombine to raise the efficiency of the pumping actions; the volumetric efficiency of the 
cylinder increases, which finally means that for the same piston displacement the power 
cylinder receives a greater quantity of cooler air than if the suction stroke had been 
carried out in the cylinder itself. The frictional resistances and the volumetric efficiencies 
of the charging pump cylinder then depend only upon the controlling conditions of the 
design, and their values should not be far different from those attainable in all good 
air pumps, say a suction pressure of from 14.1 to 14.4 lbs. absolute and a volumetric 
efficiency t) £ from .95 to .97. Since the most favorable pressure of the scavenging agent 
is probably not higher than from .75 lb. to 3 lbs. by gauge, the rise of temperature 
due to compression is negligible, and, in any case, it disappears during the transfer of 
the air to the reservoir or the power cylinder. 

In the above it is of course assumed that the scavenging air or the charge is 
furnished by a separate pump, independent of the power cylinder. The conditions 
ehange materially when, as is sometimes done, the front end of the cylinder is used as 
a pump, in which case, irrespective of constructive difficulties, the pump efficiency is 
•seriously affected by the strong heat transfer from the power cylinder to the pump 
cylinder. The case is still more unfavorable when, on account of considerations of low 
eost of machine, the enclosed crank case is made to serve as a pump. The neces¬ 
sarily bad shape of the crank case volume, the very large clearance spaces, and 
the leakage through bearings, together with the considerable heating that the charge 
experiences, combine to make the entire charging action much less efficient than in the 
case of the 4-cycle engines. This all leads to the conclusion, at least for the larger 
machines, that if it is intended to utilize to the fullest extent the advantages of the 
two-stroke cycle, only independent pumps, which may be designed as such without 
restriction, should be used. The use of crank case pumps is inefficient even in small 
engines and should be permitted only when cheapness and light weight are the primary 
requirements. 

2. The Phenomena of Scavenging, clearing the cylinder of burned gases, are in them¬ 
selves quite complex and, as already pointed out above, are by no means fully cleared 
up. The 2-cycle principle will stand or fall with the scavenging of the power cylinder. 
Efficiency, reliability, capacity—in short, everything depends upon the thoroughness 
with which the burned gases are driven out of the cylinder. If, as is usually intended, 
the cylinder volume is to be cleared out with air, the introduction of a certain amount 
of air in excess is indispensable. There must therefore be available a volume of 
scavenging air greater than the cylinder volume, because during the scavenging period 
some air is certain to be lost through the exhaust ports and, unless some excess air is 
at hand, some of the burned gases are certain to remain. In the case of independent 
pumps the available volume of air may be made anything, depending upon the pump 
dimensions; if, on the other hand, the front end of the cylinder, or even the crank 
case, is used to compress the air, an excess of air is not obtainable. There is in such 
cases likely to be a deficiency, since the otherwise possible volume is further decreased 
on account of large clearance spaces and heat transfer from the power piston. If it is 
inadvisable to use a separate pump, it then becomes necessary, by intelligent choice of 
duration of the scavenging period and proper design of the cylinder and valves, to 
obtain, with the least possible loss of air, as nearly perfect a cleaning out of the 
cylinder as the conditions will allow. The more or less perfect solution of this 


44 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


particular problem determines almost entirely the practical value of any new 2-cycle 
construction. 

As long as the scavenging medium is led directly from the pump into the main 
cylinder, its proper action is seriously hampered. With pump and main cylinder cranks 
180° apart, it is possible to scavenge only to the dead center; where the pump 
crank leads the main crank, the scavenging air must have comparatively high pressure 
on account of the short time available for transfer. It seems necessary therefore to 
interpose between pump and power cylinder a receiver of such size that, during the 
entire scavenging period, the air pressure can be maintained without much drop. This 

pressure should be very low. An abnormally high scavenging pressure not only 

increases the lost pump work, but, what is worse yet, interferes with a thorough 
driving out of the burned gases. The author, on the basis of practical experience for 
the last fifteen years, must emphatically state that the much favored increase in the 
pressure of the scavenging air entirely fails to fulfill its purpose and that a cleansing 
of the power cylinder both thorough and economic is possible only with air moderately 
compressed. Air highly compressed enters the cylinder with great velocity, and 

rebounding from the inner walls, causes eddy currents of such magnitude that, from 
the outset of the scavenging period, burned gases mix with the incoming air and thus 
a part of them is retained in the remaining air or the new charge. The perfect 

scavenging action on the other hand shows very different characteristics. The air 
should enter the cylinder slowly, avoiding all counter or eddy currents, and should, if 
possible, in the form of a solid column drive the exhaust gases in like manner ahead 
of itself out through the ports. This action, however, can only be obtained with low 
pressures and sufficiently large ports. Success in ridding the cylinder of burned gases- 

lies entirely in carefully avoiding any breaking up of the exhaust gases and in prevent¬ 

ing the intermixing of air and burned gas. 

The allowable minimum pressure of the scavenging agent depends mainly upon the 
size and frictional resistance of the transfer ports and upon the time available. 
Neither the size of the ports nor the time available can, for very obvious reasons, be 
made too great. Mathematically the question may be treated as follows: 1 t Let the 
pressure in the receiver (or pump) be p lbs. per sq.in., and the temperature be T°; 

also let the pressure in the power cylinder be po lbs. per sq.in.; then the velocity 

of the air will in general be 2 

w=<fi58^T^l—^^ ft. per sec.(67) 

Then, if the area of the ports equals / sq.ft., the volume of air transferred per second 
will be 

V = awfcn.it .(68) 

In these equations, <£ represents a velocity coefficient and a the coefficient of 
contraction; the product of these two factors is the so-called coefficient of efflux p = a(f>. 
In practice it has been found that <j> varies from .85 to .95. The value of a cannot 
be definitely given since it depends upon attendant conditions, and it should therefore 
be determined for every case. It is safe, however, in most cases to make a approxi¬ 
mately equal to from .6 to .65. 

1 A comprehensive mathematical investigation of the dynamic action during scavenging and 
charging has been published by A. Wagener in the Gasmotorentechnik, 1903. 

2 Hutte, ed. 19, Vol. I, p. 332. 





CRITICAL EXAMINATION OF THE VARIOUS CYCLIC EVENTS 45 
Example. Assume <£ = .90, a = .62, T=461+81=542°, p = 16.3 lbs., and p„ = 14.9 lbs. Then 


w = .9 X 58 \5 _ I^l) = P er SeC ‘’ 

and V — . 62X367/= 227/ cu.ft. per sec. 

It will be seen from this that even with small pressure differences the velocity of 
flow is considerable. 

The author’s requirement calling for “lowest possible pressure of the scavenging 
air ” is not accepted without qualification by Mr. A. Wagener, who for about ten 
years has been engaged in studying the dynamic actions in the cylinder of the Oechel- 
hauser engine. The reasons for this dissent are given in the following in Mr. Wagener’s 
own words: 1 

• , ‘H. Giildner, in his “Entwerfen und Berechnen der Verbrennungsmotoren, ” concerning 
air storage receivers, voices the opinion that it is desirable to work during the period of 
scavenging with as small a pressure drop as possible and hence to make the receivers as large 
as possible. This requirement seems quite justified for certain types of engines; but to generalize 
it and to make it apply without qualification to all 2-cycle engines would be a mistake, for 
the reason that for certain 2-cycle engines it is contradictory to another requirement of fully 
as great importance. Such engines are those utilizing the lean industrial gases, in which the 
air for both scavenging and charging comes from the same receiver and flows through the 
same transfer ports, and whose speed regulation must be within very close limits. When a rich 
fuel is used in an engine, the amount of such fuel required per cycle even for maximum load in 
large engines is comparatively small. It is quite possible to force this amount of gas into the 
cylinder without simultaneous addition of any considerable quantity of air and still obtain a 
sufficiently intimate mixing of the fuel with the volume of air already present in the cylinder. 
In engines of the Oechelhauser and Junkers type, for instance, the charging pump was made to 
force a certain amount of illuminating gas to which only a small amount of air had been added 
into the cylinder during the compression stroke. Under these conditions the combustion line 
testified to the formation of a very good fuel mixture. Under low loads on the engine it was 
even possible to force fuel gas alone into the power cylinder with equally satisfactory results. If, 
on the other hand, lean gases are to be used, the gas volume required for each working cycle 
is, in general, so large that a method of operation like that above outlined is entirely unsuited. 
What would happen is that only at the beginning of the gas-charging period would there be 
any mixing of the incoming gas and air, for the rest of the time the incoming gas column 
would merely displace the air. With lean gases, therefore, it becomes necessary to form the 
fuel mixture in the charging pump or the proper quantity of air must be injected into the 
power cylinder at the same time with the gas, so that the mixture may form during the 
period of introduction. Experience has shown that there are serious objections to the first of 
these schemes, that is, to form the mixture in the charging pump, one of which is that the 
receivers and passages are filled with an explosive mixture. If under these conditions, owing to 
the formation of mixtures deficient in air—a state of affairs which may be, but is not always, 
prevented, especially at starting—the charge should be ignited during the transfer period, the 
result would be the explosion of the mixture in the receiver, which accident nearly 
always results in damage to some part of the engine and especially to the charging-pump 
valves. 

If the mixture is formed by simultaneous introduction of gas and air, it becomes necessary, 
when the amount of fuel used is changed for purposes of speed regulation, to correspondingly 
change the amount of air if the combustion is to remain the most efficient possible. The 
governing arrangements, by which the problem of measuring the required quantities of gas and 

1 From a lecture delivered before the Bavarian District Society of German Engineers, 1903. This 
lecture with some elaborations was published in the Gasmotorentechnik, and deserves attention on account 
.of its mathematical treatment of the charging and discharging actions of 2-cycle engines. 



46 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


air must be carried out, may be constructed with the most simple means, and with promise of 
greatest reliability, when, in order to control the volumes of the gases transferred, the pressures 
in the receivers are suitably changed. The most economic way of changing the pressures is by 
decreasing the volumes of the gases delivered by the pumps, so that with a decrease in pump 

capacity there shall be a corresponding decrease in the pump work. The simplest means of 

obtaining this result is probably to arrange a by-pass valve under the control of the governor, 
so that at the beginning of the discharge stroke a larger or smaller quantity of gas or air, 
depending upon the position of the valve, is forced back in the suction mains at practically 
atmospheric pressure. 

Now, however, if the change in the quantity of gas transferred is to be brought about by a 
change in the pressure in the receivers, it becomes desirable to make the volumes of the latter, 
including that for air, as small as possible. The reason for this obviously is that, the greater 

the volume of these receivers, the greater will be the time elapsed between the moment that 

the governor begins to affect the gas volumes delivered by the pump and the moment that the 
volumes delivered from the receiver begin to be affected in the corresponding way. 

A drop in pressure of say 1.5 lbs. would call for a receiver volume equal to approximately 
210 cu.ft. With such volumes, governing arrangements as above outlined would fail completely; 
and the .question therefore arises whether the requirements that the air receiver should be made 
as large as possible, is of enough importance to warrant seriously curtailing freedom of choice 
regarding the details of construction. 

In Oechelhauser machines results satisfactory in every respect have been obtained with 
dimensions of the gas and air receivers such that, for the maximum gas volumes transferred, 
the pressure drop is as high as 14 lbs. per sq.in. This construction calls for comparatively high 
pressures at the beginning of the scavenging period, up to 22 or 23 lbs. per sq.in. absolute for 
high speed engines. In spite of this fact the difficulties mentioned in Guldner’s book have not 
manifested themselves up to this time in Oechelhauser engines. It is of course quite possible 
that the ring of ports used for the air inlet in this machine is specially suited for such condi¬ 
tions of operation. Giildner says, among other things, concerning this question: “Air highly 
compressed enters the cylinder with great velocity, and rebounding from the inner walls, causes 
eddy currents of such magnitude that, from the outset of the scavenging period, the burned 
gases mix with the incoming air and thus a part of them is retained in the remaining air or the 
new charge. This is no doubt practically what happens when the air is introduced from one 
side of the cylinder only, as for instance, when the ports are confined to only a part of the cir¬ 
cumference. If, on the other hand, the inlet ports are arranged completely around the circum¬ 
ference, the writer is of the opinion that, in view of the initial guidance that the air currents 
receive when they strike the face of . the piston and their subsequent mutual interference, the 
final result is the formation of a column of air of sufficient solidity to clear out the cylinder. 
In practice, the minimum volume of scavenging air required to obtain reliable operation serves 
in every case as a criterion for the efficiency of the scavenging action. Whether the introduction 
of this minimum quantity of scavenging air is always combined with the greatest economy of 
operation is altogether another question. 

It might finally be mentioned that the magnitude of the pressure drop has but very little 
influence upon the work required by the pumps. If, for instance, in any given case the pressure 
drop is 8.5 lbs. per sq.in., the charging pump must of course compress to about 23 lbs. per sq.in. 
which is comparatively high; but the transfer from the pump to the receiver commences almost 
with the. beginning of the compression stroke, since, owing to the previous operation, the 
pressure in the receiver is practically atmospheric. On the other hand, if the pressure drop is 
say 1.5 lbs., the pump has to compress to only about 19 lbs.; but since during the previous 
operation the receiver pressure has dropped to only 17.5 lbs., the pump must first compress to 
this pressure before the transfer to the receiver can begin. Comparing the indicator cards from 
two pumps operating under these conditions, it will be found that the work required in the 
former case is but little greater than that in the latter. The most important factor determining 
the work required. by the pump is the value of the mean receiver pressure p m , but the latter 
has no direct relation to the volume of the receiver.” 

Another requirement for good scavenging is that the air be not admitted until the 
exhaust is complete, that is, until pressure equilibrium in the cylinder has been 
established. An earlier admission of air is possible only with unnecessarily high air 


CRITICAL EXAMINATION OF THE VARIOUS CYCLE EVENTS 


47 


pressure, the use of which, as already shown, is bad in itself and should be avoided. 
The burned gases are still urider high pressure and consequently their temperature is 
higher than ordinary, which tends to decrease the weight of air or new charge in the 
cylinder owing to the heating they receive. 

It is quite difficult to attain a complete pressure drop to atmosphere, or nearly so, 
with the ordinary exhaust valve in the short time available in most cases. A ring of 
ports in the cylinder wall is much better suited to the purpose, and consequently much 
used in the design of 2-cycle engines. But even in this case it is best not to admit the 
scavenging air until the dead center, to make certain under any circumstances that the 
pressure drop is complete and that the exhaust gases are cooled as far as possible. 
With the piston controlling the exhaust ports, there will then still be available a period 
of time for scavenging and charging equal to the time required by the piston to cover 
10-12% of its stroke at that end of the cylinder. Better results will thus be attained 
than if the air had been forced into 'the as yet highly compressed and heated gases at 
an earlier period. As compared with 4-cycle machines, the beginning of compression in 
2-cycle machines may commence a little later on the return stroke without much 
harm, because in the former the compression must commence with a pressure p a 



Fig. 24. Fig. 25. 


invariably somewhat below atmosphere, while in the 2-cycle engine the compression 
commences with a pressure of from .5 to 1.0 lb. per sq.in. above atmosphere 
(see p. 52). 

It is desirable to have the pressure drop in the air receiver as small as possible, in 
order that the scavenging may continue wHh undiminished force to the end of the 
period. The best means of attaining this end is to make the receiver as large as 
possible and to commence the transfer of air not too early in the stroke. The two 
diagrams of Figs. 24 and 25 make the points clear. The first is taken from a receiver 
of too small volume. The pump delivers air into the receiver from a to b. From b to 
c the pressure remains constant at about .3 atm. (4.2 lbs. per sq.in.) above atmosphere. 
Scavenging commences at c, but at the outer dead center d the pressure has already 
dropped to .1 atm. (1.4 lbs. per sq.in.) and beyond this point the drop is very slow. 
It is evident that the transfer of air had already ceased bef<?re the inlet valve closed at 
o; that is the best part of the time available for scavenging has not been used. After 
a considerable increase in the volume of the receiver, diagram Fig. 25 resulted. The 
maximum scavenging pressure is now a little less than before; the transfer commences 
a little later, but continues at about the same rate until the closure of the inlet valve, 
since in this case the receiver pressure does not drop below .2 atm. (2.8 lbs per sq.m.) 
above atmosphere. The pressure at the beginning of compression in the power cylinder 
is correspondingly higher, and since on account of somewhat later introduction the air 
is less highly pre-heated, tfyere will be a greater charge volume in the cylinder, which 
finally means greater engine capacity. 

The manner of scavenging without the use of intermediate receivers is well 
illustrated in Figs. 26 and 27, the diagrams being taken from a crank-case 2-cycle 
engine of the author’s own design of 1895. The crank-case pump draws air from a 
to 6, Fig. 26, and compresses it into the large clearance spaces up to the point c, 










48 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 




-approximately to .3 atm. (4.2 lbs. per sq.in.). After the burned gases have been 
■exhausted from the cylinder, the transfer of air proceeds with rapid drop of pressure. 
Scavenging is completed at the point d, because the piston then commences its return 
stroke, and the little overpressure still remaining in the crank case suddenly disappears. 
In order to make the scavenging period independent of the movement of the piston, 
the transfer pipe between crank case and cylinder was made as large as possible, and 
was separated from the crank case by a flap check valve. In this way the pipe was 
converted into a receiver, although of insufficient capacity. The diagram obtained, 
Fig. 27, sufficiently illustrates the improvements thus made in the scavenging process. 
The check valve admits the air to the transfer pipe, but prevents its return, and thus 
permits a continuation of the scavenging action even after the crank-case pump has 
started on its suction stroke. This is about the only satisfactory way by which the 
burned gases can be removed from the cylinder in the case of engines using the crank 
case as a pump. 

The rapid and complete drop in the pressure of the burned gases to near atmos¬ 
phere after the exhaust valve has opened, is of extreme importance in case the fuel 
mixture instead of air is used for scavenging. Even in this case a little air is usually 
introduced ahead of the mixture, but this does not in itself offer safety. It is even 
possible that, owing to this introduction, some of the unburned gas particles contained 


c 

- 



(A ---- 

■C _ 

a. 

Fig. 26. 

: 2 

V: i 

Scavenging 

Fig. 27. 


in the burned gases may be ignited by the fresh air, thus firing the entire incoming 
charge upon entrance. The only safeguard against an occurrence of this kind is to hold 
hack the scavenging medium until the exhaust of the burned gases is complete. 

The difficulties that may be encountered through neglect of the above rule are well 
indicated by Mr. A. Wagener in a lecture delivered before the Society of German 
Engineers at their general convention in 1900. Mr. Wagener tells of some of the 
experiences encountered in putting into operation a 600 H.P. Oechelhauser and Junkers 
blast-furnace gas engine as follows: 1 


An attempt to operate the machine at normal speed failed because, even after a comparatively 
short period of operation at less than normal load, the working cylinders became so hot that 
the occurrence of pre-ignition was feared, and the latter actually did occur from time to time 
in the right-hand cylinder, as during the tests this machine probably carried somewhat more load 
than the other half of the unit. 

The trouble in this case was found to be caused by the fact that the exhaust ports did 
not open early enough, and that on the opening of the air ports a part of the burned gases 
escaped through these into the air jacket surrounding the air ports. This action could be 
traced at a speed as low as 120 r.p.m., although the interference was not strong enough at this 
point to cause serious trouble. 

The inter-related actions of air and burned gas are very clearly shown in Fig 28 The 
upper curve (1) shows the pressure of the burned gases in the cylinder during exhaust.' The 
line was taken by means of a weak spring, and transferring it to the drafting-board the pressure 
ordinates were further enlarged with the greatest care. At the same time that line (1) was 
obtained a pressure diagram (2) was obtained from the air jacket surrounding the air ports, 
^f drawn + 0ver t0 the same scale. In obtaining this line, the reducing motion 

of the indicator was set parallel with the outside cranks, so that the abscissas represent the 


1 Z. d. V. D. Ing., 1900, p. 1517. 









CRITICAL EXAMINATION OF THE VARIOUS CYCLIC EVENTS 


49 


positions of the back piston. Fig. 28 also shows the relative proportions and positions 
of the air, mixture, and exhaust ports. The exhaust ports are of course controlled by 
the other piston, but have been transferred in the figure with due regard to the relation 
between stroke and port opening. In this way, if an ordinate be drawn at any point 
corresponding to a given piston position, that part of the areas not cross-hatched in the figure, 
lying to the right of this ordinate, will represent the port opening for each ring of ports at that 
piston position. 

The exhaust ports open at c, but at first the pressure drop is not very marked. This is probably 
partly due to friction in the passages, and also to the fact 
that in any case a certain amount of time is required to 
produce the necessary acceleration even in the comparatively 
small mass of the burned gases. Further, there is some lag in 
the propagation of the pressure drop from the neighborhood of 
the exhaust ports to the middle of the cylinder where the 
indicator is located. The same interesting phenomenon 
occurs in the pressure diagram for the air jacket (2). The 
air ports start to open at a, at a point where the pres¬ 
sure in the cylinder is still 47.3 lbs. per sq.in. abs. One 
would assume that the immediate action would be for the 
exhaust gases to rush into the air jacket, raising the pres¬ 
sure in the latter considerably. It will be seen, however, 
from the diagram that this occurs at a point considerably 
later in the stroke. At the point d the pressure in the 
' cylinder is the same as that in the air jacket, and it 
would naturally be expected that from here on both lines 
should indicate an approximately similar pressure drop. The 
diagram, on the other hand, shows a further increase of 
pressure in the air jacket, and a continued drop in the 
cylinder. 

The further path of curve (2) shows that soon after the 
pressure in the air jacket attains its maximum value, the mix¬ 
ture, consisting of burned gases and air, rushes into the cylinder 

with such high velocity that a vacuum is formed in the air jacket itself. This in turn is 
followed by a rise in pressure due to the air flowing in from the main receiver. Near the dead 
center, therefore, very complicated inter-actions occur in a very short space of time, and the 
masses of gas and air are set into violent vibration, a phenomenon to which the writer will 
again refer later on. 

The diagrams of course only show pressure variations in the vicinity of the indicator, and 



\ Mixture 

'wm 

Wm 

Air 

wmmmimm 

1 



Fig. 28. 





it should not be forgotten that the lines are apt to be somewhat modified by the inertia of the 
moving parts of the indicators themselves. 

It is very interesting to determine the rapidity with which the actions described occur. From 
the moment of opening of the air ports, for instance, to the moment that the exhaust gases 
commence to enter the air jacket with rapidity, that is, from a to b, the time interval is 
.0054 sec., during which the pressure in the cylinder drops from 47.3 lbs. abs to 38.0 lbs. abs. 
If it be assumed that the pressure drop is uniform throughout the volume of the cylinder, it 



































50 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


follows that in this short time 54 ltr. (1.906 cu.ft.) of burned gases escape through 
the ports. The mean port area is 257 sq.cm. (42.7 sq.in.), and hence, if the gases flow 
through the ports at approximately atmospheric pressure, the mean velocity would have 
to be about 365 m. per sec. (1197 ft. per sec.). It is of course nearly impossible to determine 
the absolute velocity, since no reliable means are at hand for measuring the pressure differences 
between the inlet to and exit from the ports, but the above computation, even if very 
approximate, shows that the maximum velocity of the outflowing burned gases reaches an 
extraordinarily high value. 

The difficulty above outlined was remedied in the machine under discussion by elongating 
the exhaust ports toward the middle of the cylinder, a distance of 20 mm. (.8"). 

Fig. 29 shows the complete diagram from the air jacket as taken before changing the 
exhaust ports, while Fig. 30 shows the diagram as at present obtained. It will be seen that 
the sudden pressure drop and the violent vibrations of gas and air are almost entirely eliminated. 
Further, no vacuum is formed at any time between the air ports and the controlling valve on 
the air receiver, and the entire operation of transferring the air takes place under much quieter 
and more efficient conditions. 

The machine now operates without difficulty at 130 to 135 r.p.m., and, even under overloads, 
the cylinders heat up to so slight an extent that the hand may be laid for a few moments upon 
the uncooled outer ends of the cylinder without discomfort. 


III. Comparison between 4= and 2=cycle Engines 

1. Considered from the standpoint of theoretical thermal efficiency, both principles 
of operation possess exactly the same value, their working cycles do not differ in any 
respect. The real thermal efficiency, however, is affected by a series of practical 
conditions, several of which speak very strongly for the 2-cycle engine. Since the specific 
capacity of the 2-cycle engine is twice that of a 4-cycle machine, it follows that the 
former requires only one-half of the cylinder volume of the latter. The comparatively 
greater superficial area of the smaller cylinder volume is in favor of the 2-cycle engine 
in so far as relatively more heat of compression is transferred to the cooling water, 
permitting the use of higher compression pressures. Besides possessing practical advan¬ 
tages, the latter fact means a better thermal efficiency. It is a fact of course that the 
greater heat transfer, on account of relatively greater superficial area, also exists during 
combustion as well as compression. This is not a desirable condition, but it seems that 
the disadvantage is about balanced by the fact that the average combustion tempera¬ 
ture of lean mixtures highly compressed is correspondingly less. It has been found 
that the losses to cooling water in the two types of engine are not far different, in 
fact this loss sometimes shows less in 2-cycle machines than in 4-cycle machines of 
like capacity. A 15 H.P. Benier 2-cycle engine, for instance, tested by Witz, required 
61 lbs. of cooling water per B.H.P. hour, while an Otto engine of the same capacity, 
examined by Kohler, required 54 lbs., both being operated with producer gas. A 600 
B.H.P. Simplex 4-cycle blast-furnace gas engine, according to Hubert’s tests, required as 
much as 110 to 143 lbs. of jacket water per B.H.P. hour. From the author’s own 
experiments, small 2-cycle engines lose to the cooling water from 3400 to 3800 B.T.U. 
per I.H.P. hour, the larger sizes from 3000 to 3300 B.T.U. As in the case of 4-cycle 
machines, this amounts to from 35-40% of the heat supplied. 

Fear might be expressed that the scavenging air may become strongly pre¬ 
heated by the burned gases to be displaced. This however does not take place if the 
scavenging is properly done, i.e., if the transfer of air does not commence too soon. 
The temperature T a of the charge at the beginning of compression in 2-cycle machines 


CRITICAL EXAMINATION OF THE VARIOUS CYCLIC EVENTS 


51 


is usually not higher, rather lower, than in 4-cycle engines. 1 This fact may be easily 
explained by considering that the transfer of air, a process in itself combined with a 
temperature drop, takes place with great rapidity, taking up from about ^ to | of 
the time of one stroke, while in the case of the 4-cycle the charging action occupies 
the time of one entrie stroke, hence the heat absorption by the charge must be much 
more marked. A rise in the temperature of the charge after the closure of the transfer 
ports no longer effects the engine capacity.' The latter is increased by the fact that, in 
2-cycle engines, compression commences when the charge is at a pressure already 
slightly above atmosphere, hence the charge weight is greater than if the charge had 
been drawn into the cylinder by suction. Assuming that at the end of the scavenging 
period the pressure in the cylinder of the 2-cycle engine is say 1.05 atm., while the 
pressure at the end of the suction stroke in the 4-cycle cylinder is say .9 atm.; then, 
irrespective of any influence that the charge temperature may have, the 2-cycle 

cylinder contains — X100 = 16% more charge weight than the 4-cycle cylinder, 

and has hence gained the same amount in specific capacity. (For the reason that the 
external useful work is, other conditions being the same, directly proportional to the 
charge weight.) 

2. The second important basis of comparison of the two engine types is found in 
the friction losses of the machine. In this respect the 2-cycle engine is under a 

. disadvantage as compared with the 4-cycle in that its charging and discharging actions 
require a greater movement of the volumes of the gases involved, since the charge 
must pass through two cylinders. All other conditions being the same, this fact would 
require a greater amount of work, and this is what is usually found in ,2-cycle engines 
as constructed to-day. It should, however, be noted that a charging pump specially 
constructed for its work can operate at a higher efficiency % as a pump than the main 
power cylinder, and also that in the 2-cycle engine the entire exhaust stroke is 

eliminated, i.e., the resistance p r does not exist. Estimating the suction pressure in the 
4-cycle diagram at —.1 atm., the resistance during exhaust at +.1 atm., at the same 
time taking the vacuum during the suction stroke of the 2-cycle pump equal to 
.03 atm. (which should be easily reached), the gain in the 2-cycle engine would be 
.l + .l -.03 = .17 atm., and this amount, with skilful and intelligent design, should be 
sufficient to cover the extra work required in the transfer of the charge in the 2-cycle 
engine. That the moving of greater volumes of gas may not represent an increased 
loss of work is quite clearly shown in some tests by Humphrey on a Crossley and a 
Premier engine, both large horizontal 4-cycle machines (see Part III). In spite of the 
fact that in the Premier engine the cylinder was very thoroughly scavenged (air excess 
250%) the work done in charging and discharging at the greatest total load of 650 

I.H.P. was only 5.5%, while the same work in the Crossley engine, of appoximately 

the same capacity, represented 7.3% of the total load. 2 Consequently, although at 
present the pump work in 2-cycle engines is still from 8—10% of the I.H.P., as com¬ 
pared with 6-7% in 4-cycle engines, it does not necessarily follow that this defect is 


1 The comparatively stronger heat transfer during compression and the lower temperature T a 
at the beginning of compression in 2-cycle machines made themselves strongly felt in the first 
Diesel-Guldner 2-cycle engine by the fact that the compression had to be higher than in 4-cycle 
engines by from 60 to 90 lbs. in order to produce an end temperature T c sufficiently high to insure 
regularity of ignition. 

2 See Engineering, 1901, p. 197. 




52 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


permanent; the further development of 2-cycle types will undoubtedly yet wipe out 
the 2 or 3% of difference still existing. 

The greater part of the friction loss of the machine is found in the rubbing parts. 
With regard to this no definite comparison can be made, since no practical data is 
available, and mathematical determinations of the friction losses based on arbitrary 
assumption would be worse than useless. Some light may, however, be obtained from 
a few general considerations. 

According to an old reliable rule of thumb, the friction loss in similar piston engines 
is the greater the greater the volume of piston displacement per horse-power. According 
to this rule, 4-cycle and 2-cycle engines would then be on an equal footing if the volume 
of the pump be added to the cylinder volume. When it is considered, however, that 
the 2-cycle pump piston works against very low pressure only, that therefore it may be 
fitted with very light rings and be an easy fit in the bore, and further that this light 
piston operates all the time on cool, well lubricated surfaces, there can be little doubt 
as to which cycle possesses least pump friction. When a piston, built for operation 
against pressures from 300 to 400 lbs. per sq.in., is for one-half of the time called upon 
to operate against pressures of from 3-4 lbs. per sq.in., i.e., pressures averaging only 
1% of those for which the machine member was designed, there must evidently be 
useless sacrifice of mechanical efficiency and of economy in material. Such, however, is 
exactly the condition of things in the 4-cycle engine. 

It is not at all necessary that the scavenging pump should be an independent 
machine, an added complication requiring a certain amount of work for its operation. 
It is possible, for instance, to so design the cross-head as to make it act as an air 
pump, as was done in the Diesel-Guldner engine. This construction secures the added 
advantage that the power piston and cylinder are relieved of the side thrust of the 
connecting rod. (It is pointed out in Part III that to make the power piston a 
machine member, already fully loaded by the explosion pressure and working on highly 
heated surfaces, also carry considerable side thrust, is fundamentally a mistake.) The 
adoption of such special designs makes it possible to construct an engine with pump, 
in which the piston and guide friction is not only not greater, but even smaller, than 
in an engine without the pump. That this can be done is shown by the example of 
the Benier engine already quoted, this engine with scavenging pump showing a 
mechanical efficiency of 81% at only two thirds load. The 500 H.P. Premier engine 
above mentioned showed a mechanical efficiency of 88.8% without the pump, and 83.8%, 
allowing for the pump work, while the Crossley engine quoted at maximum load 
showed a mechanical efficiency of only 83.0%. 

As compared with the work of charging and discharging and the friction in pistons 
and cross-heads, the rest of the friction losses are of secondary importance. It need 
only be briefly pointed out that, since the piston of the 2-cycle engine has approxi¬ 
mately only one-half the area of the 4-cycle piston for the same capacity, the crank 
mechanism and shaft may be made correspondingly smaller, and the friction loss in 
pins and bearings is correspondingly less. The latter is further decreased by the fact 
that the fly-wheel of the 2-cycle engine is lighter by approximately 50%. This is 
corroborated by the common practical experience that the mechanical efficiency of a 
given engine drops a few per cent, when an extra heavy wheel (for electric light 
service) is substituted for the ordinary wheel. 

3. From the standpoint of design, the relation between the 4-cycle and the 2-cycle 
engine has already been made sufficiently clear in the preceding articles. The superiority 


CRITICAL EXAMINATION OF THE VARIOUS CYCLIC EVENTS 


53 


of the latter, as regards regularity of operation, space required, weight, and cost of 
construction, is patent in all cases where the proper design has been employed. In the 
building of large engines, the complexity of the 4-cycle construction more strongly 
emphasizes the advantages of the 2-cycle, especially so the nearer the sizes approach 
the limit of capacity. For the large machines the main issue of the “cycle” controversy 
is no longer the fuel consumption, but, very simply, the possibility of construction 
under ordinary manufacturing conditions, and of reliability of operation. The fact that 
beyond a certain size the 4-cycle engine fails to meet these conditions in the best way, 
has served to clearly show the advantages of the 2-cycle machine, and has done much 
to give it its new lease of life. Even Ernst Korting, the successful builder of 4-cycle 
engines, agrees with the writer in this, stating on one occasion 1 that “the successful 
construction of 2-cycle engines, either single- or double-acting, is almost a necessity in 
the development of large units, such as required in the operation of steel works. The 
weight of the various machine parts and their dimensions increase so enormously, when 
it is attempted to generate these powers in one 4-cycle cylinder, and, further, the 
growing complexity of the machine and the unreliability of operation become so serious 
when it is attempted to get the same power by the combination of four single-acting 
4-cycle cylinders, that either type of construction can only be considered as a make¬ 
shift until an efficient and reliable 2-cycle machine appears on the market” 

Practically the same conditions exist to-day, although the builders of large 4-cycle 
engines have again taken up the construction of double-acting engines, thus doubling 
the specific horse-power capacity. (In this connection it should be noted that this 
method does not increase the capacity of the cycle itself, but, by making a given 
cylinder double-acting, merely raises the capacity of that cylinder.) A double-acting 
4-cycle machine is on the same footing regarding capacity as a single-acting 2-cycle 
engine; but since, previous to the construction of these double-acting 4-cycle engines, 
large 2-cycle engines had already been built, the relative positions of the two types of 
machines are not changed in the least. Concerning the question of regulation, i.e., the 
uniformity of tangential effort at the crank, the adoption of the double-acting principle 
in the 4-cycle improves things but little as compared to the 2-cycle. The regulation of 
the double-acting 4-cycle engine compared to that of the single-acting 2-cycle engine is 
still in the ratio of 2 to 3, in favor of the latter machine, and in order to create 
anything like equality in this respect when the 2-cycle is made double-acting, two 
double-acting 4-cycle cylinders would have to be used. 

4. Concerning the question of economy, there is little more to be said, after pointing 
out the practical equality of the two principles of operation regarding the thermal 
efficiency and friction losses. It can easily be seen from tests on the 2-cycle engines of 
Benier, Guidber, and Diesel-Guldner that in every case the indicated thermal efficiency 
T)i is very satisfactory, and that, even in cases where the machine is unduly handi¬ 
capped by friction losses in the pumps, the economic efficiency y w in 2-cycle machines 
is but little inferior to that found in good 4-cycle engines. Considering the question of 
economic operation in its entirety, i.e., not only considering fuel economy but also first 
cost, maintenance, reliability, etc., any 2-cycle engine of fair grade but of absolute 
reliability will be quite able to take care of itself. 

The requirement laid down by Prof. Riedler, that for best results in the building 
and operation of large gas engines the cycle should be the two-stroke cycle with 


1 Z. d. Y. D. I., 1902, p. 127. 


54 METHODS OF OPERATING GAS ENGINES AND GAS ENGINE CYCLES 


capacities up to 500 H.P. in the cylinder, 1 at first seemed to many to be impossible of 
realization. But the question has in the meantime grown in economic importance and 
in a short time the final solution will most likely not depend upon what can be done, 
but upon what must be done. 

1 This is Riedler’s opinion as laid down in “ Schnellbetrieb ” (Part XI, p. 37), three years ago. Since 
that time Prof. Riedler has developed into a very strong advocate of the 4-cycle engine, in fact, of a 
certain double-acting 4-cycle engine. When the author in spite of this adheres to Prof. Riedler’s early 
favorable opinion of the 2-cycle principle, he does this because it seems to him that in the meantime 
conditions affecting the problem have not changed in the least, and also because facts there cited by 
Prof. Riedler in his usual clear manner seem to him to point much more strongly to the conclusion 
above quoted than to the diametrically opposite view expressed before the 45th General Convention of 
the Society of German Engineers. 

The expression coined by Prof. Riedler on this occasion, that 

“the gas engine will revert to its 
original starting point, that is to the 
4-cycle principle,” 

certainly does not offer any serious obstacle to the further development of the 2-cycle engine. It will 
probably temporarily aid certain 4-cycle constructions—it will not, however, serve to maintain the 
4-cycle principle in the large engine field. Qui vivra verra! 


PART II 


THE DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


A. FUNDAMENTAL CONSIDERATIONS 

I. Less “Invention,” more “Rational” Design 

The most serious handicap to the development of any type of machine is an 
aimless multiplicity of its various constructed forms. These many .forms, not called 
forth by any real need but mostly through the ambitions of restless inventors, lead to 
designs which, in nine cases out of ten, are at variance with even the most fundamental 
principles of ’construction. And although such performances soon find their proper 
reward in brevity of life, there still remains this general disadvantage, that the endless 
procession of such one-day constructions has a deterrent effect upon sane efforts in the 
branch of industry concerned, and that all constructions, whether capable of life or 
not, are weighed down with the same appearance of unreliability. 

This abnormal tendency to invention was especially pronounced in the case of 
gas-engine construction. There was a time—and it is not even now entirely passed 
during which the employment of known and tried devices was frowned upon, when not 
only 'the uninitiated but also the professional men were strongly attracted by what 
seemed new or unknown, and when the happy inventor of a hot tube or an exhaust 
muffler could at a turn of the wrist become the founder and director of a gas-engine 
works. It went without saying that every “independent ” designer, besides being 
-conversant with the principles of ordinary design, had to have a new valve gear or a 
new method of regulation of his own, and it has actually happened that one of these 
inventive geniuses (?) designed a fly-wheel to act at the same time as a tank for 
cooling water, merely to show something “never before offered.” No wonder, therefore, 
that the practical value of most of the new engine systems was in the inverse ratio to 
the number of patents covering them. 

If anything has contributed more than any other cause to the internal soundness 
of our present day gas-engine practice, it is the fact that this reckless inventing, this 
search after new things, is gradually being abandoned. In its place we find business¬ 
like design and development of those types which have managed to maintain them¬ 
selves through the disorder of earlier years, and which have become, to a greater or 
less degree, fundamental forms. For the gas-engine designer of to-day the most 
important problem is to develop these fundamental forms with the aid of the well- 



56 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


known elements of general machine design, and upon the basis of principles proven by 
science and tested by practice. In the solution of this problem, the less free rein is 
given to the inventive bent, the better for the designer and the problem. This applies 
especially to the younger designer whose ambition is very apt to lead him to look 
upon inventing as a necessary preliminary to, or even as the ultimate aim of, his 
labors. In doing so he loses sight of the real aim of every sane engineering problem, 
i.e., to furnish something which embodies the most advanced and best ideas, and which 
shall be of practical utility. We do not design in order to invent: on the contrary, for 
important engineering work it is usually one of the requirements to its successful 
accomplishment that it be carried out only on known and tried ground. With a great 
deal of reason, therefore, a good designer generally has the name of being an unsuc¬ 
cessful inventor—and vice versa. 

It is of course not in the least the intention to underrate the importance of 
inventions which are based upon the recognition of an actual need, and which have 
their origin in true mental labor; but how little of this finds expression in the 
ordinary run of patents covering designs and constructions! 

Here as everywhere in first place comes the alpha and the omega of every theory of 
design, which is the proper adaptation of the constructive means at hand to the existing 
conditions. 

That which is unquestionably the right thing in one case may be entirely out of 
place in another. A small engine must be designed from an entirely different stand¬ 
point than is proper for a large industrial machine; a marine engine calls for a 
construction differing from that of a stationary motor; and an originally stationary 
oil engine is by no means a locomobile, because it is placed upon four wheels. Only 
in one sense is there for all forms and types the same requirement, that is, reliability 
and durability, and to it all other requirements must be subordinated. A single hour 
of stoppage in operation may serve to wipe out the saving in fuel for an entire year, 
and an engine showing 50% thermal efficiency would be a costly prime mover, if after 
a few years it should have to be relegated to the scrap pile. 

In general, the following points should be kept in mind: In the case of the small 
engine, greatest possible simplicity, low cost, and durability; for the large engine, 
greatest economy, and for the automobile machine, smallest possible weight. Compared 
with the primary requirement, as determined in any given case, the rest are of 
secondary importance, or do not enter the problem at all. Of course, in the case of 
the small engine, fuel economy should be considered, just as we should take simplicity 
of construction into account in the case of the large machine, but the primary 
requisite should not be lost sight of in each case. There is no sense in doubling the 
number of machine parts and the manufacturing cost of a commercial machine to gain 
a saving of a few cubic feet of gas; but it is just as unwise to hesitate over the 
addition of another valve, or of a few pounds of metal, when the reliability and the 
economy of the machine can be increased thereby. 

The working stresses in internal-combustion engines are from two to three times as 
high as those found in steam-engine practice. Besides this these stresses cannot be as 
certainly controlled in their time of occurrence as can those of a steam engine, and it 
may occasionally happen, due to a combination of circumstances, that they are changed 
in a manner detrimental to the entire construction. This fact makes it imperative that 
the designer choose only such forms in which all maximum stresses occurring are taken 
up as centrally as possible; which admit of definite strength computations and which 
preclude any hidden defects in manufacture. Plain forms are not only statically the 


FUNDAMENTAL CONSIDERATIONS 


57 


safest and cheapest as regards manufacturing costs, but also pleasing to the technical 
sense if properly carried out. A tendency to excessive economy of metal, however, is 
apt to degenerate this plainness of form into mere ugliness, and this should be guarded 
against. It is for instance out of place to give the frame of a machine the shape of a 
plain box with beveled edges to save the cost of a better frame pattern, as is done in 
many English and American machines. Since every frame pattern is used a large 
number of times, its cost cannot seriously add to the manufacturing cost of a 
machine. 

The good features of even the best design are not well brought out when the 
arrangement of the auxiliaries and the painting of the machine is out of keeping with 
the rest. An engine which, against a bright-red background as a body color, shows a 
sample assortment of nickeled parts, highly polished brass, and artistically scraped cast- 
iron surfaces, etc., may be an attractive exhibit for the masses, but to the professional 
man its appearance is repellant. The finished machine should show two colors only: 
a dull black or, at least, dark color, for the cast surfaces and the metallic white for 
the machined parts. The nickeling of entire machine parts is technically wrong and 
perhaps allowable only in the case of small auxiliary parts of brass, bronze, or copper. 
The external surfaces of machined parts need only be neatly polished or ground; high 
polish is out of place in machine construction, and makes an unnatural impression even 
in engines “on parade.” The more perfect the design, the better the materials and 
the more accurate the construction, the less need is there for any embellishment—true 
worth speaks for itself. 

In the general arrangement of the installation the qualifications and knowledge of 
the attendants must not be left out of account. It is best not to form too high an 
estimate of them, even in the case of small engines. An attempt to explain a 
complicated installation by a lengthy set of operating directions invariably fails in its 
purpose; the directions are either not read, or, if read, are wrongly interpreted. But 
few machine parts and fewer motions on the part of the attendant should be required 
to start and stop the machine and to keep it in regular operation. Single parts, which 
during the assembling of the machines are accurately set, and especially such parts in 
which change in position by inexperienced hands can become dangerous to the engine, 
had better be permanently fastened to place them out of the reach of attendants 
anxious to investigate or to improve. 

Another point which should receive the proper attention of the designer is the 
item of repair, by which is meant the readiness with which worn or broken parts can 
be repaired or replaced by the ordinary machinist. In this respect it is of course not 
necessary to go as far as a certain landowner who expected that a serviceable oil- 
locomobile should be “capable of repair with a hoe”; but those parts at least which 
are sure to wear, as for instance, bearings, hot tubes, packings, studs, and bolts, etc., 
should be so designed that in case of need they can be renewed in every good 
machine shop. That important parts, like pistons and valves, should be replaced by 
the factory only, is self-evident. 

Finally, the manufacturing cost of every new construction must not be forgotten, 
because this determines the sale price, which in turn affects the marketability and the 
commercial economy of the machine. The influence of commercial economy is very 
often not clearly understood. As is well known, the real economy of a machine is 
determined by the total operating expenses, among which interest on the investment 
and depreciation form a large, and often the largest, part. When, therefore, as an 
example, one machine costs $100 a year less for fuel than another, but costs $200 more 


,58 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


in interest and depreciation, etc., evidently the first engine is not a profitable invest¬ 
ment in spite of its high thermal efficiency. 


II. Horizontal or Vertical Type 

The horizontal type of engine has maintained the field since the early days, and it 
is usually assumed as proven that the vertical type is adapted to small sizes only. 
That is not true; the author is rather of the opinion that the vertical type of engine, 
with the crank-shaft below the cylinder, is also of great utility for the larger sizes, and 
is, in many important points, superior to the horizontal form. A few lines will show 
this. From steam-engine practice we know that vertical engines always show a some¬ 
what better mechanical efficiency than the horizontal, and this is corroborated by the 
results on gas engines. Further, the losses through leaky pistons are always considerably 
less in the vertical than in the horizontal types. The reasons for this are obvious. 
The lost work due to the weight of pistons and cross-heads of horizontal machines is 
eliminated in the vertical type. In the latter the piston moves concentrically in the 
cylinder bore, with the allowable play equally distributed, and for every piston position 
the rings can make good contact without undue restraint. In this way the packing is 
made more perfect, and the piston friction, together with the amount of wear, are 
reduced, especially since in the vertical engine cylinder lubrication is more uniform and 
certain. The bending stresses occurring in the w r alls of the frame of horizontal 
machines, which can hardly be eliminated, are easily avoided in vertical engines. In 
the latter type the maximum pressure in the cylinder can be transmitted in an 
exactly central line to the crank, and the frame, thus protected against bending 
stresses, can be made correspondingly lighter. This point grows in importance as the 
piston pressures to be transmitted increase, and that explains w'hy distinctly high- 
pressure engines, like the Diesel and the Banki, adhere so persistently to the vertical 
form. 

Besides this, the vertical type of construction gives a form of cylinder which, from 
the standpoints of manufacturer and of efficiency of combustion, is the most favorable, 
because all valves can be placed in the cylinder cover, and the combustion chamber 
can be kept free from all ports and dead spaces. The number of cylinders can without 
trouble be increased to give any desired power capacity, while at the same time a 
nearly complete balancing of the moving parts is possible, thus allowing of increased 
speeds of rotation. Since to-day internal-combustion engines are as yet mostly con¬ 
structed without cross-head, the relation of the cylinder diameter D to the stroke S 
is not of as great importance here as it is in the case of the steam engine, and for 
that reason large gas engines with long stroke, or with slow speed of rotation, can be 
built without requiring excessive head-room. The foundations of vertical machines are 
affected in the main by vertical forces only. They can therefore be made smaller or 
cheaper than those for horizontal machines of the same power, and in some cases may 
even be dispensed with altogether. The taking out and replacing of pistons for 
cleaning purposes, etc., is easily accomplished even in vertical machines of the greatest 
capacity. Valves and valve gearing also are free and easily accessible, and are not, as 
is the case in large horizontal machines, hidden in the foundation or in trenches. 
Suction, exhaust, and water pipes can be put up in the most suitable manner and in 
the most direct w r ay without the necessity of ditching, etc., at the place of erection. 
Short lengths of suction pipe, in easily accessible position, are of especial benefit in the 


FUNDAMENTAL CONSIDERATIONS 


59 


case of suction-gas plants. Such pipes facilitate the removal of all dusty or tarry 
deposits, which in the case of horizontal machines, is often difficult, thus giving a 
better guarantee for reliability of operation. As in the cost of erection, the vertical 
engine is also cheaper in cost of manufacture and in maintenance and attendance. If 
besides this, we take into consideration that a vertical motor is free to expand axially, 
due to the heat generated in operation, can be most easily direct-connected to any 
consumer of power, and takes only about two-thirds of the floor space of a horizontal 
machine of the same power, we shall have touched upon all of the important advan¬ 
tages of this type of construction. 

There are, however, cases in which the horizontal type has the preference. As 
long as we do not possess gas-cleaning apparatus, which is both perfect and economical, 
a large blast-furnace gas engine for instance should be horizontal. The reason for 
this is that any mechanical impurities in the gas are apt to adhere to the piston faces 
and cylinder bores of vertical machines, and may easily do damage. In horizontal 
machines, on the contrary, the combustion chamber and the valve ports may be so 
designed that the solid impurities of the gas can be swept out to a great extent by 
the exhaust gases, or can, at least, be kept away from sliding surfaces. What is said 
here regarding blast-furnace gas applies more or less to all industrial gases and 
especially to the coal-dust engine. Again, for double-acting engines the vertical is not 
the best form, because the- placing of the valves in the lower head and their operation 
in this position would cause trouble. The lower head itself could be but imperfectly 
water-cooled. For very large power capacities the vertical type of engine is out of 
the question, as it is also in the case where pump or air cylinders are to be connected 
to the extended piston rod. 

Very often the statement is made that the valve gear of vertical machines is more 
complex and involved than that of the horizontal. The writer does not consider this 
well founded, and will show later that it is possible to provide the vertical engine with 
a valve gear of simple and mechanically correct form. Further, it is stated that the 
crank-shaft of vertical engines is too much concealed, that it is hard to watch and to 
keep in order. To meet this it is merely necessary to point to the vertical machines 
of great power with which our war-ships and our merchant marine are equipped. 
These machines are subject to long-continued service operation under conditions much 
more severe than those found on land. A properly designed and well constructed 
■crank-pin should not require anxious care if it is kept cool, protected from dust, and 
well lubricated. All of this, however, is accompiished with more ease in a vertical 
than in a horizontal machine; in the first place because the bearings are less subject 
to the radiant heat of the cylinder; secondly, because the frame body can be tightly 
enclosed; and, lastly, because the shaft so enclosed can be abundantly oiled without 
los 3 of lubricant and without the fouling due to flying grease. Even a complete 
enclosing of the frame does not affect the ventilation of the crank case if the interior 
is connected to the outer air by means of an equalization pipe. The engine room, 
however, will by this means be kept free from the often bothersome radiation of heat 
and the vapor from overheated oil. 

According to all of the above, there should be little doubt that up to 150—200 
I.H.P. per cylinder the vertical type of engine is, in general, fully the equal of the 
horizontal, and is in fact .in many points superior. It is therefore to be expected that 
ihe vertical engine will have a future as a commercial gas power machine. 


60 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


III. With or Without Cross=head 

When, a few years after the appearance of Otto’s gas engine, internal-combustion 
engines were no longer constructed with a cross-head, the maximum explosion pressures 
were from 140-170 lbs. per sq.in., while the mean pressure of the expansion rarely 
exceeded 50 lbs. per sq.in. At that time gas-engine practice was dominated by the 
wishes and requirements of the small power user, who, more strongly then than to-day, 
was looking for lowest prices, minimum room requirement, and greatest simplicity. In 
the meantime the small engine grew to be a first-class prime mover in the commercial 
field. The working pressures increased to from 350-500 lbs. per sq.in., the mean effective 
pressure to 110 lbs. per sq.in. and over. Such changed conditions naturally demanded 
much stricter requirements in the construction of this type of heat engine, and it 
would seem to the writer that among these requirements the question “With or 
without cross-head” deserves special consideration. 

Nobody would look upon a steam engine without a cross-head as quite or entirely 
first class; on the contrary, stationary steam engines with trunk, pistons have acquired 
so bad a reputation that they have one after the other disappeared from the market, 
or, at least, have been forced to a position of little importance. It is certainly true 
that one experience can perhaps not be generalized, and that these opinions of the 
designer of steam engines need not be considered binding by the constructor of gas 
engines. But a very small part of the mistrust, founded on the facts of hard experience, 
that the former entertains for the entire absence of cross-head, would not come amiss ta 
the latter. The writer will try to prove this. 

The omission of the cross-head in the case of small and even medium-sized gas 
engines, say up to 125-150 I.H.P. per cylinder, is without doubt allowable; the trunk 
pistons can in these cases be made so long that the unit normal pressures on the 
piston barrel and the cylinder wall are very small, thus affecting durability but 
little. Although, in this construction, cylinder lubrication requires greater attention and 
the piston rings are apt to blow through quicker than in case a cross-head is used,, 
these disadvantages seldom cause trouble in commercial machines, and they are in 
most cases counterbalanced by the advantages of this method of construction. In the 
case of large engines, however, the conditions are materially different. If, in these 
machines, the connecting rod thrust is to be kept within limits safe for pistons and 
cylinder walls, and if this lateral pressure is not to interfere seriously with the work 
of the piston rings, the trunk piston must be made of such abnormal length as ta 
cause trouble both in manufacture and in operation. With the size of piston we find 
an increase also in the internal and the expansion stresses, and in the inertia forces. 
The problems of properly fitting the sliding surfaces and of their lubrication finally 
present insurmountable difficulties. To realize this it is only necessary to remember 
that piston lengths exceeding 6 ft. have been proposed. How is it possible for the 
designer, the shop, or the operator, to absolutely control machine elements which are 
of such extreme size and at the same time of such vital importance? Which method 
of cylinder lubrication is capable of satisfactorily and economically oiling a piston 
surface of 60-80 sq.ft.? The union of several important operations or offices in one 
machine member is always a weak point, but it is entirely bad when these various 
offices are to each other as fire and water, i.e., when they are naturally contradictory. 
The most appropriate example of this practice is the trunk piston of some of our 
single-acting large gas engines. The purpose of the piston is in the first place to act as 


FUNDAMENTAL CONSIDERATIONS 


61 


a machine element transmitting gas pressure without leakage, and although it is just 
barely able to perform this primary duty, it has besides this also been utilized as a 
heavily loaded machine member taking up lateral thrusts, much to the detriment of the 
proper performance of its chief office. That was a mistake. The piston should be 
designed to its smallest detail purely as an element to transmit gas pressure without 
leakage losses, the same should be done for the cross-head as an element to resist 
lateral thrust, and the two should not be united. The fact that the separate cross¬ 
head increases the length of the machine by the length of the piston rod and that the 
manufacturing cost is increased by a few per cent, is fully counterbalanced by better 
construction, less liability to repair, and greater life of the machine. 

An important advantage of this construction is the single fact that the piston pin, 
winch is usually hard to get at, difficult to keep in order, and often of barely sufficient 
size, is replaced by the easily accessible cross-head pin, which can be made with any 
desired dimensions and is easily lubricated and taken care of. 

The use of a separate cross-head has no noticeable effect upon the lost work of 
the machine, i.e., the mechanical efficiency is not affected. This is proved by results 
obtained, from Otto and Diesel engines. In the case of large engines even the contrary 
may be expected, because the coefficient of friction for a properly designed cross-head is 
evidently less than that for piston and cylinder surfaces highly heated, imperfectly 
lubricated, and subject to varying expansions and contractions. Under conditions so 
unfavorable to good lubrication it is really self-evident that the internal rubbing 
surfaces must be loaded as little as possible and consequently that they should be at 
the outset relieved of any lateral thrusts. The latter no longer amounts to a few 
hundred pounds only, in some of the largest engines it exceeds twelve and even fifteen 
tons during the expansion stroke, and it seems to the writer that it would pay to 
take care of such forces of transmission in the proper place. 

As a necessary part of double-acting engines, and not on the basis of the above 
considerations, the cross-head again finds extended application in recent gas-engine 
practice. For the time being, this important step in advance is of benefit only to a 
few large industrial types, but it is beyond question that in the near future all large 
■gas-engine design will revert to the use of the separate cross-head. 


IV. Single= or Double=acting Cylinders 

The first gas engines were double-acting and great consumers of gas and oil, 
which in a few years led to the abandonment of the use of the front end of the 
cylinder in order to keep the piston and cylinder bore cool by free convection. A 
little later this led to a considerable simplification of the entire construction in that 
cross-head and piston rod were also done away with. Following this, for a number of 
years, the double-acting principle was looked upon as impractical. For small and 
medium sized machines this view is probably quite justified, because to build them 
double-acting would make them more complex without good reason, would increase the 
length of the machine and raise the manufacturing cost. In such a case a two-cylinder 
machine is on all accounts cheaper than a double-acting single cylinder. 

The case is different, however, for large machines. The complete development of 
this type of machine and special considerations regarding its field of application 
<lemand the greatest possible cylinder capacity, and this is best obtained economically 
by decreasing the piston displacement required per unit of power, that is, through a 


62 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


smaller number of working strokes and the use of both sides of the piston. In 
consequence of the more uniform crank efforts of double-acting machines, the weights 
of rotating parts can be made less than for single-acting engines (for the 4-cycle, for 
instance, the decrease is from 35^0%). 

The great importance of this for large powers will be shown on page 63. A 
further practical advantage of double-acting machines is seen in the fact that the 
main driving mechanism of the engine is mechanically better utilized on account of 
greater specific power capacity. This, combined with the elimination of lateral thrust 
on the piston, causes a comparatively smaller amount of lost work, which in turn 
means higher fuel economy per brake horse-power for the same thermal efficiency at the 
cylinder. 

These considerations have in the last few years led designers to again take up the 
construction of double-acting gas engines, and through the use of this principle several 
new and promising types have already been developed. The results so far obtained 
seem also to point strongly to a more general use of this method of operation for 
medium sized engines. 

The employment of both sides of the piston makes it necessary, as in the case of 
large trunk pistons, to water-cool the piston and usually also the piston rod, and no 
great difficulty is encountered in doing it. The question of the stuffing-box, which was 
formerly so important, can to-day be considered solved for all practical purposes. As 
a result of experiments on the older types of single-acting tandem engines, and after 
years of practical experience, serviceable metallic packings for piston rods have been 
developed. The most difficult part of the design is found in the proper shaping of the 
cylinder heads, because the water-cooled stuffing-box leaves but little room for valves- 
and igniters, and the number of openings in the head is apt to render it weak and 
less free to expand and contract. It is especially difficult to satisfactorily design the 
crank-end cylinder cover. In small engines this difficulty is not so strongly felt, 
because in that case the valve openings may be placed in the sides of the cylinder, 
thus keeping the head free and incidentally simplifying the valve gear. For large 
machines, however, such design is not satisfactory, for constructive and practical 
reasons, and should not be used, if only for the reason that the compression space is 
apt to be badly cut up. 

In several of the more recent designs of large double-acting gas engines, however, 
the cylinder heads have been simplified by placing the inlet and outlet valves at the 
ends of the cylinder in a perpendicular plane through the cylinder axis, as is very 
often done in poppet-valve steam engines. It is true that by this means the cylinder 
heads are rendered much more satisfactory; further, it becomes easier to get at the 
inside of the cylinder and at the piston; on the other hand the constructive and 
operative difficulties before found in the heads are now transferred to the cylinder 
itself, i.e.^ to the most important part of the entire machine. Thermally, also, this 
valve arrangement is not satisfactory because it involves, even more so than in the 
case of the old form of cylinder heads, a form of combustion chamber not favorable to 
good combustion of the charge. 1 There is little possibility that this difficulty can ever 
be entirely overcome, and it is likely, therefore, that this type of large double-acting, 
engine will labor under a certain thermodynamic disadvantage for some time to come. 
Practically, however, this point has little weight when, in all other respects, the double¬ 
acting type has the advantage over the single-acting. 

1 Translator’s Note. The objections to this type of cylinder seem to have been quite successfully 
overcome, as it is to-day used in many large engines. See Part IV. 


FUNDAMENTAL CONSIDERATIONS 


63 


This condition of things, i.e., complicated cylinder heads and unsatisfactory 
combustion chambers, can be very much relieved by employing the piston to operate 
the exhaust through a ring of ports at the middle of the cylinder. In such a case one 
valve is done away with and room is gained. Exhaust ports operated by the piston, 
however, can only be used in 2 -cycle machines, which leads to the idea of combining 
the double-acting with the 2-cycle principle, as has already been done in the Korting 
engine. Another step in advance could be made by designing the cross-head, which in 
double-acting engines is a necessary part, as a pump for the air required for scavenging, 
the cylinder, thus eliminating a separate pump. The writer first employed such a 
cross-head for a single-acting 2-cycle machine in the Diesel-Guldner engine. The 
doubling of the pump piston area required to scavenge both ends of the cylinder does 
not give any sensible trouble. Finally, the double-acting 2-cycle engine offers another 
advantage of importance in the case of large engines, in that the inertia of the 
reciprocating parts is taken Up in the compression of the charge at the end of each 
stroke, and does not cause the disagreeable reversal of pressure common to all 4 -cyclo 
machines. In a combination of the 2-cycle principle with double-acting operation we thus 
possess the most natural means of bringing the internal combustion-engine in an economical 
form to its highest capacity. 


V. Multi=cylinder Arrangements 

The reason which determines the employment of several cylinders irv the case of 
the steam engine, i.e., further decrease of fuel consumption, does not exist for the gas 
engine. In the case of the latter, considerations of construction and manufacture, in 
connection with manufacturing costs, compel a division of power between several 
cylinders. Let us first consider the 4-cycle engine. Although the explosion pressure 
occurs every fourth stroke only, and even then decreases very rapidly, it is necessary 
to build the driving mechanism, and in fact the entire machine, to meet the highest 
pressure. On that account the diameters of the shaft journals and especially the crank 
pin become comparatively large, which in turn means greater friction losses and which, 
in the case of large cylinder diameters, may lead to unwieldy and very costly construc¬ 
tions. This useless massing of metal, together with the increasing lost work, is the more 
strongly felt when we consider that the mean pressure of the 4-cycle strokes is only a 
comparatively small part of the explosion pressure (see p. 68 ). The influence of this- 
unfavorable pressure ratio upon the driving mechanism of the engine is also strongly 
manifested in the weight of rotating masses required to insure the necessary uniformity 
of rotation. A single-acting 4-cycle engine requires a fly-wheel weight of at least 
110 lbs. per brake horse-power when the peripheral speed of wheel is about 66 ft. per 

second and the coefficient of regulation as high as ^ = For a 500 H.P. machine of 

this type, therefore, the fly-wheel would weigh about 55000 lbs., which is roughly one- 

third of the entire engine weight. For electric lighting operation, with £=^r, the 

engine at the same rotative speed would have to have a wheel weighing about 

187 000 lbs.; and for the running of alternators in parallel, with <5 = ^ 5 ’ even a wei ght 

of 330 000 lbs. would be required. Such enormous rotative masses increase the cost of 
manufacture in a manner nearly prohibitive, on the one hand, owing to their own cost. 


64 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


on the other to the necessary increase in the size of shafts, bearings, foundations, etc. 
Thus it happens that, above a certain engine capacity and a certain coefficient of 
regulation, two-cylinder and multi-cylinder engines can be built at less cost than single¬ 
cylinder machines of the same capacity, owing to the fact that on account of much 
more uniform crank efforts much lighter fly-wheels can be used. 

Not less serious are the practical difficulties which oppose the increase of capacity 
of large single-cylinder 4-cycle engines. Parts of the machine, enormous alike in their 
dimensions and their weight, require in their manufacture unusually heavy machine 
tools and other costly shop facilities which cannot be operated economically because 
they are not steadily employed and are used only occasionally to their fullest capacity. 
For this reason, together with the fact that the manufacture of the important parts is 
complicated and often uncertain, the manufacturing costs become inadmissibly high and 
periods of delivery are greatly lengthened. 

In the finished machine new difficulties are brought to light. Cylinders and 
cylinder covers, which, merely on account of their size and complicated form, are not 
easily kept free from internal stresses during the course of manufacture, are subjected 
to additional stresses, owing to uneven expansion under the heat developed in opera¬ 
tion. This may easily cause a total stress exceeding the elastic limit of the material 
-and may eventually lead to warping and rupture of the parts concerned (see pp. 120 
and 131). Many thousand parts have already been lost in this way. The exact fitting 
of the unwieldy pistons in the cylinder bores becomes extremely difficult; too much 
play means constant leakage, i.e., loss of gas and of pressure, noisy operation, etc.; 
too tight a fit causes large friction losses through piston friction and consequent 
scoring of the cylinder walls. These losses are aggravated by difficulties which operate 
against certain and uniform lubrication of the large piston surfaces. In a limited 
degree water-cooling of the piston offers a useful remedy for these defects, but the 
gain is bought at the expense of an undesirable increase in the number of machine 
parts and an increase in the weight of the reciprocating masses. 

Although such unfavorable mechanical conditions are in themselves sufficient to 
seriously affect the economy and reliability of large single-cylinder 4-cycle machines, a 
further heavy disadvantage comes in when we consider that the thermal operations 
inside of the cylinder also lose efficiency slowly as the cylinder sizes increase beyond a 
certain limit. Since the ratio of superficial area of the combustion chamber to its 
volume grows smaller as the stroke volumes increase, and the walls are made thicker 
from considerations of strength, less heat will be conducted through the walls and more 
will be stored up in them, which means that the allowable compression pressures must 
be steadily decreased (see p. 52). Consequently the thermal efficiency and the capacity 
per unit stroke volume will be correspondingly decreased, and the difficulty already 
encountered in satisfactorily igniting and completely burning such large charge volumes 
is still further magnified by the increase of compression volumes. This point is 
especially serious because these large machines are ordinarily run on lean gases difficult 
to ignite. That is the reason why in single-cylinder engines the heat supplied is 
thermally best utilized in sizes between 50-100 H.P., and why only in isolated cases is 
there a gain shown up to 150 H.P. Beyond this capacity we usually find a loss 
rather than a gain in economy. 

Led by these considerations, German designers to-day seldom go beyond 300 B.H.P. 
in a single-cylinder single-acting 4-cycle machine operating on lean industrial gases. 
Greater capacities are obtained either by decreasing the number of strokes to a cycle 
or, if the 4-cycle is retained, by making it double-acting or by increasing the number 


FUNDAMENTAL CONSIDERATIONS 


65 


of cylinders. Under such conditions the work can be handled -with facility, both in the 
shop and the drafting room, and the machines possess a greater degree of reliability, 
since in case of need every cylinder, or each side of a piston, can be operated 
independently. 

The lower limit for the division of engine capacity among several cylinders 
depends mainly upon the requirements of regulation; and, in the case of auto-engines, 
upon the required degree of balance; for the rest it is left to the judgment of the 
designer. In the case of stationary single-acting horizontal engines, two cylinders are 
usually not employed until the capacity exceeds 100-150 H.P.; for stationary vertical 
machines, however, the limit of division is usually lower than this. For auto-engines 
the deciding elements in this respect are mainly satisfactory counterbalance of the 
moving parts and the room available for installation. 

Multi-cylinder arrangement, as applied to the construction of large units, has lost 
considerably in importance since the extended introduction of the double-acting type. 
After a short period of trial in this field these three- and four-crank constructions are 
considered quite as obsolete as the old so-called opposed engines, and they conse¬ 
quently find application to-day only in special designs, as for automobile machines, 
etc. In stationary practice the aim is to employ two cylinders at most, and these may 
be arranged either in tandem or side by side (twin engine) with the fly-wheel between 
them. Only the largest sizes, exceeding 1000 H.P., compel the use of four cylinders, 
which are usually arranged as two tandem engines side by side. 

The comparative values of the usual multiple cylinder arrangements, with respect to 
the variation of crank efforts or the regulation, are given numerically in Tables 22 and 
23, pp. 230 and 231. From these it is seen that the single-cylinder single-acting 
4-cycle machine with ignitions 720° apart shows the least favorable result, while the 
double-acting two-cylinder 2-cycle or the double-tandem double-acting 4-cycle with 
ignitions every 90° of crank angle show the best. Referred to the weight of fly-wheel 
rim to be furnished per unit of power to two the extreme types of engine above 
mentioned, for otherwise exactly similar conditions, the ratio is found to be about 
65 to 1. This fact natrually becomes most important in the case of large engine 
capacities and rigid requirements of regulation. 

In general, the double-acting tandem machine has the disadvantage that the inside 
of the cylinders and the pistons are not easily accessible. Within certain limits this 
may be remedied by making the connecting piece between the two cylinders of 
-sufficient length and of simple open form, although this means a further increase of 
the already great length of this type of machine and a decrease in the stiffness of the 
connecting member. On the other hand, the tandem machine can usually be built 
somewhat cheaper than a two-cylinder (side by side, so-called twin) machine of the 
same capacity. Another advantage of the former is that the second cylinder may be 
installed later, as the demand for power may require, and by doing this the coefficient 
of regulation for the original fly-wheel may be improved from 20 to 25% in spite of 
the doubling of engine capacity. 

The weight of the reciprocating parts is relatively considerably greater in a tandem 
engine than in a twin machine, and consequently the rotative speed of the former is 
limited. From the standpoint of reliability of operation the tandem machine is also a 
little inferior to the twin, because any accident to the driving mechanism of the 
former, as for instance the heating of the crank pin, shuts down the entire machine, 
while in the case of the twin engine one cylinder may be kept in operation. 

The foundations for tandem engines are always somewhat less simple than those 


66 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


for twin engines, but the difficulties here encountered can be overcome when the 
machine parts concerned have been properly built and the ground under the foundation 
is safe. 


VI. Complete Expansion and Compounding 

If the expansion volume of a charge is the same as the suction volume, that is, 
when the ratio of expansion is equal to the ratio of compression, which is practically 
always the case in engines fully loaded, the burned gases will, at the moment of 
exhaust, be under a pressure of from 35 to 60 lbs. which is not utilized. This loss may 
be diminished if the burned gases are allowed to expand beyond the suction volume, 
which may be done, for instance, by making the effective suction stroke shorter than 
the expansion stroke. This problem has occupied designers from the time of Otto, and 
some of the older solutions of it are found in the gas engines of Charon. The simplest 
solution was given by Kohler, 1 who, by closing the inlet valve before the end of the 
suction stroke, makes the charge volume taken in less than the piston displacement. 
The Korting Brothers first put this idea into practice in 1891 in their “precision ” 
gas engines, but perfected it by putting the time of closure of the inlet valve under 
the control of the governor, thus suiting it to the load. Although this re.sults in a 
very flexible system of speed regulation, the real aim of Kohler’s idea, an increase in 
the thermal efficiency, was not thereby realized. The reasons for this are: At full load, 
with this method of regulation the suction volume is nearly or quite equal to the 
expansion volume; with a decreasing load the expansion becomes more and more 
complete, but at the same time the compression pressure decreases, and any advantage 
gained by the former is nearly or entirely lost by the unfavorable influence of the 
latter. 

If complete expansion is to show economic gain, the compression pressure must 
remain normal for all loads; in such a case the indicated thermal efficiency is really 
the greater, the further such expansion is carried, p. 12. But the economic efficiency 
does not increase in the same ratio. Every increase in the expansion volume entails a 
decrease in the mean working pressure, and consequently also a lowering of engine 
capacity per unit volume. From this it follows that the engine mechanism is less 
completely utilized, and that the ratio of lost work to indicated work becomes greater. 
Thus, at some point, the theoretical advantage of complete expansion is changed to a 
practical disadvantage. How far expansion can be carried with economy cannot in 
general be definitely stated, but it is certain that the lower the normal compression 
pressure, the richer the mixture, p. 37, the slower the combustion of the charge, and 
the smaller the engine friction, the greater may be the ratio of expansion. Anything 
less than a final expansion pressure of 15 lbs. by gauge at full load is rarely used, 
because under such conditions the mean piston pressure or the capacity per unit volume 
would become too small. 

In conformity with the practice of compounding in steam-engine design, attempts 
have from the first been repeatedly made to utilize the expansive force of the exhaust 
gases in a second cylinder, but always .with equal failure. At best the constructed 
compound gas engines showed the economy of fair single-cylinder engines, but as regards 
simplicity, cheapness of construction, and reliability of operation, they were always inferior 
to the latter. The reasons for the failures of these experiments are to be found mainly 


‘Kohler, “Theorie der Gasmotoren,” Leipzig, 1877, p. 22. 


FUNDAMENTAL CONSIDERATIONS 


67 


in the. high temperatures encountered and in the low specific heat of the burned gases. 
To obtain in the low-pressure cylinder a mean effective pressure at all commensurate 
with the friction loss, it is necessary to cut short the expansion in the first or 
combustion cylinder. The gases, still under high pressure and often as yet in process 
of combustion, then have a temperature far exceeding 1800° F., and for that reason 
the valves between the cylinders can be kept in fair shape only by thorough water¬ 
cooling. But this deprives the gases of a considerable proportion of their quantity of 
heat, and this, together with fluid friction losses, causes a decided drop in pressure. 
The loss of heat during the transfer of the gas is incredibly high. The writer remem¬ 
bers a case in which the cooling water from the two intermediate valves carried away 
as much heat as was transformed into indicated work, and for which the actual 
indicator diagrams showed only one-half the area of the ideal. It is also found that 
the jacket loss from the high-pressure cylinder is very large because, owing to the early 
exhaust opening, the mean temperature of the cylinder is considerably higher than that 
found in single-cylinder engines. This high mean temperature in turn has an unfavor¬ 
able effect upon the lubrication of the high-pressure piston, especially since the. mean 
pressure is also far above the normal. 

Taking it altogether, experiences so far hardly admit of a doubt that from the 
use of extended or complete expansion we can only expect an economic gain in the 
case of machines employing low compression, and even in such case only when the 
idea is carried out in one cylinder and with the simplest means. Compounding is 
and will remain without promise in gas-engine construction. 


VII. Ratio of Stroke to Diameter and the Speed of Rotation 


In theory purely thermal considerations appear to make it advantageous to build 
gas engines with shortest possible stroke and the highest possible rotative speed, in 
order to shorten the time of combustion and expansion to the greatest possible extent, 
and thus to decrease the heat losses to the cooling water. For cooling is a function of 
its time of action, and this is decreased when, with the same piston speed, the 
stroke $ 

ra ti° c p ameter * s sma U- But for any machine, practical and not theoretical conside¬ 


rations have the call, and in this case practice makes demands directly opposed to 
theory, i.e., greatest possible stroke and moderate speed of rotation. 

The reasons for this are various. In the first place it should be considered that a 
decrease in the length of the stroke means a corresponding decrease in the time 
available for the various operations of charging and discharging; the mixture is apt to 
be less uniform and the combustion less complete; the volumetric efficiency of the 
cylinder falls off with a consequent decrease in the charge volume. It should not be 
forgotten that for half of its time of operation the gas engine is an air pump and all 
conditions which affect the efficiency of such a pump affect the internal-combustion 
engine in a like manner. It is known that the volumetric efficiency of the suction 
stroke is, in general, but little dependent upon the rotative speed, if the stroke is long, 
as compared with cylinder diameter. This makes it possible for the designer to choose 

S 

high speeds of rotation with high ratios of as long as he properly adapts the mixing 


and ignition arrangements to the high speeds. The more perfect the mixture of air 
and of fuel, the purer the mixture itself, and the more efficient the ignition apparatus 



68 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


the higher may be the piston speed—provided that thereby the efficiency of the 
machine as a pump is not decreased. This may be easily prevented by the use of 
correctly designed and properly actuated inlet and outlet valves. The largest stationary 
gas engines of to-dajr already employ piston speeds exceeding 780 ft. per minute 

with a ratio of ^ = 1.2 to 1.3. This present limit is certainly capable of extension if the 

s 

ratio is taken at a more favorable, i.e., a higher figure, and if by the expedient of 

making the cylinders double-acting the momentum of the reciprocating engine parts 
can be cushioned against the compression. 1 

The length of the stroke also has an important influence upon the shape of the 
combustion chamber, whose volume is of course proportional to the stroke volume. 
The present high compression pressures already make it difficult to so place inlet and 
outlet valves of sufficient size in the cylinder heads that enough internal cooling 
surface remains. The unfavorable effect of port openings upon the amount of cooling 
surface and upon the strength and elasticity of the casting itself is the more seriously 
felt the shorter the length of the compression space. In the case of cylinder heads of 
long-stroke machines, more space is available, it is easy to properly place the valves, 
and the port area amounts to a smaller part of the total enveloping surface of the 
combustion chamber. 

To this must be added the fact that, for equal piston displacement, that is, for 
equal engine capacity, the longer the stroke the smaller will be the cylinder diameter, 
and consequently also the maximum distance of flame propagation. Since the length of 
the compression chamber is in all cases only a part of the cylinder diameter, an 
increase of the former at the expense of the diameter is of advantage to ignition and 
combustion. 

The error in the use of x sliort strokes in gas engines is, however, most strongly brought 
out when we examine the effect upon engine proportions. It is clear that the strength 
of every machine part must be computed on the basis of total piston pressure resulting 
from the maximum pressure of explosion. The capacity of the machine, however, does 
not -depend upon this total maximum but upon the total mean pressure, and the metal 
in all parts of the machine is therefore better utilized the nearer maximum and mean 
pressures are together. If the mean pressure is comparatively low, the aim must be to 
also decrease the maximum pressures, which, with a given piston displacement or engine 
capacity, can only be done by decreasing the cylinder diameter and increasing the 
stroke. Now in the case of 4-cycle engines the ratio of mean to maximum working 
pressure is very low (with an explosion pressure of 350 lbs. and a mean effective 

92 

pressure of 92 lbs. per sq.in., the ratio would be 350-r-— = 15 to 1), and therefore the 

4 

rule to obtain the necessary piston displacement with small piston area and long stroke 
applies most strongly to this type of engine. If this rule is neglected we arrive at 
machine dimensions which are uneconomical both on constructive and practical grounds. 
This becomes especially noticeable in the driving gear, more particularly in the crank 
pin, because at this point the entire force transmitted to the frame is concentrated. 
Diagrams, Figs. 31 and 32, show these conditions clearly. In Fig. 31 are given the 


‘The smaller automobile engines have already exceeded piston speeds of 1170 ft. per min.; as an 
example, the 12 H.P. double cylinder auto-engine of Remington makes 800 turns per minute, which with 
4.02" cyl. diameter and 10" stroke, gives a piston speed of 22.3 ft. per sec. = 1338 ft. per min. 


FUNDAMENTAL CONSIDERATIONS 


m 


crank-pin diameters for gas engines [as computed from eq. (33w), p. 166, with a maximum 
explosion pressure of p c = 350 lbs. per sq.in.] and for steam engines (initial steam pressure 
170 lbs. per. sq.in.) for various cylinder diameters. It is seen at once that the gas 
engine is at a considerable disadvantage. The lower curve of Fig. 32 shows the increase 

s 

in crank-pin diameter for various ratios of but for equal piston displacement or 

equal engine capacity. Now it would of course be possible, if it is desired to use short 
strokes, to increase the number of revolutions instead of increasing the cylinder 
diameter, thus leaving the maximum total pressure the same. What that tends to, 
however, is indicated by the upper curve of Fig. 32, which shows the increase in peri¬ 
pheral velocity, v, at the surface of the crank pin for constant crank pin and cylinder 

cr 

diameters, and for constant piston speeds, but for decreasing ratios , and consequently 
increasing speeds of rotation. Since the work lost in friction at the crank pin (expressed 



Fig. 31.—Wrist-pin Diameters for Gas and 
Steam Engines. 



Fig. 32. —Wrist-pin Diameters and Peripheral Speeds 
in Gas Engines for Varying Ratios of Stroke to 
Diameter. 


by the product k-v ) must not, for obvious reasons, exceed a certain amount, an 
increase in the peripheral velocity, v, calls for a corresponding decrease in the pressure, 
k, per unit of area. The unit pressure, however, can only be decreased by making the 
crank pin longer, which, in turn, calls for a larger crank-pin diameter from considera¬ 
tions of strength. Thus it is quite evident that an increased number of turns cannot 
be made to equalize the constructive disadvantages of short stroke. 

The superiority of long strokes with reference to static strength conditions is, 
however, by no means confined to the short period of explosion. It continues through¬ 
out the expansion period, the maximum torsional moment being smaller than in short- 
stroke machines of equal piston displacement in spite of the greater crank radius. 
Assume, for instance, that the piston displacement required for a given engine capacity 
is 2.65 cu.ft., and that this displacement is obtained in two engines of the following 
dimensions: 


q 

Case I. Cyl. diameter, 11.8"; stroke, 29.8"; ratio ^ = 2.5. 

Case II. Cyl. diameter, 16.1"; stroke, 16.1"; ratio ^ = 1. 
We shall then have the following results: 









































70 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Case I. 

Max. total pressure on piston 39 000 lbs. 
Pressure on piston for max. 

turning moment. 25 300 lbs. 

Max. tangential pressure at 

crank pin. 20 200 lbs. 

Max. turning moment. 298 000 in.-lbs. 


Case II. 

72 500 lbs., assuming 356 lbs. expl. pressure. 

47 300 lbs. at about 12% of stroke, or 
crank angle of 40°. 

37800 lbs. . 

307 000 in.-lbs. 


This table clearly shows that in Case II the machine parts have to be made 
considerably heavier to withstand the greater loads, although the engines are of 
about the same horse-power capacity. 

We thus come to the final conclusion that the use of ratios of stroke, S, to cylinder 
diameter, D, as large as possible is of general advantage in internal-combustion engines, and 
should be resorted to much more extensively than it is, especially when explosion 
pressures are high. The fact that long strokes require a somewhat greater length of 
machine is outweighed by the actual advantages derived, at least for stationary 
engines. 

Petroleum engines are not excluded from this rule. It is usual to justify the use 
of short strokes in this type of engine by claiming that small piston travel and high 
speed of rotation give less chance for condensation of the kerosene vapor on the 
cylinder walls. In this connection one is apt to overlook the fact that in short-stroku 
engines the amount of kerosene drawn in is smaller (on account of lower volumetric 
efficiency) and that the ratio of cylinder surface to charge volume increases. Oil 
engines especially should therefore combine longest possible stroke with highest possible 
speed of rotation. 


VIII. The Standard Indicator Diagram and the Allowable Stresses in Materials 

It would be entirely too early, and uneconomical in construction, at this time, to 
adopt in general a maximum explosion pressure of from 500-560 lbs. in strength 
computations, although this has, on page 16, been shown to be the pressure limit up to 
which gain in efficiency may be expected. As yet there are only a few types of engines 
in which the pressure exceeds 360 lbs., while the greater majority show explosion 
pressures that do not greatly exceed 300 lbs. In spite of these facts, however, it is not 
advisable in new designs to use the common figure of 300 lbs. as a basis for compu¬ 
tation. It is better to make p z equal to at least 360 lbs., and even then to adopt 
allowable unit stresses low enough to allow explosion pressures of say 425 lbs. in case 
of need without endangering the construction. In this manner a free way is left open 
for any changes in the working diagram that may appear desirable at a later date, an 
early scrapping of the patterns is guarded against and their use is less restricted. 

In the following strength computations the fundamental general formulae and 
equations are first developed and from these, special equations (whose numbers are 
marked with the letter n) applying to the standard diagram, Fig. 33, are then derived. 
This standard diagram and the special formulae based upon it simplify computations 
very much, allow of quick and certain insight into the principal interdependent 
features, and guard the inexperienced designer against faulty judgment. With a 
compression pressure p c equal to 128 lbs., and a mean effective pressure of 96.5 lbs. 
per sq.in. (most favorable assumption) the diagram shows a maximum pressure of 












FUNDAMENTAL CONSIDERATIONS 


71 


356 lbs., which is taken as the basis in strength computations. With this unit pressure 
the maximum total pressure on the piston for a cylinder diameter of D inches will be 

P 2 = 356 X.785 D 2 = 279.4 D 2 ~ 280 D 2 lbs. 

With the aid of this equation the special formulae above mentioned are derived. 
For the reasons already mentioned, it is not advisable to employ a lower total pressure 
in the computations, even if the regular use of a lower explosion pressure is intended. 
On the other hand, the special formulae are no longer applicable as soon as the explo¬ 
sion pressure exceeds 356 lbs. The outside limit of application of these equations may 
be taken at 425 lbs., provided both material and workmanship are excellent. For 
further details of the standard diagram, see page 155. 

The inertia pressures due to acceleration and retardation of the moving machine 
parts relieve the compression and explosion 
pressures on the driving gear by a few 
pounds per square inch of piston; but the 
amount of this reduction depends upon the 
velocity. Thus, for instance, at starting 
it is almost nothing, and it should not 
therefore be taken into consideration in 
strength computations. If in spite of this 
it should be taken into account, the work¬ 
ing parts may be subjected to the highest 
stresses just at the time when the work¬ 
ing pressures are under least control and 
accidental excess pressures most likely to 
occur. 

All working parts intended for engines 
using different kinds of fuel should be 
constructed for the working medium or 
fuel furnishing the maximum pressure, be¬ 
cause the differences in the explosion pres¬ 
sures are not great enough to call for a 

special construction for each combustible. The gas-engine designer in any case must 
figure with a greater margin of safety than, for instance, the designer of steam engines. 
Our heat engine is not supplied, like the steam engine, with a working medium already 
prepared and under definite maximum pressure conditions; instead it is called upon to 
develop its own working pressure by means of a complex chemical process and 
in less than a second. It is consequently unavoidable that the assumed normal 

explosion pressure may in certain unforeseen instances (too early or too late ignition, 
formation of explosive wave due to too rich mixture, insufficient cooling, etc.) be 

temporarily far exceeded. Just such “supernatural forces,” however, furnish the proof 
for conscientious design and construction of an internal-combustion engine; many a 

motor too economically designed has under the stress of such circumstances cost its 
designer very dear. 

How high the allowable stresses may be chosen for the various working 
parts is given in each case as accurately as possible. Many examples based on 
successful constructions serve to show the accuracy of these assumptions. If various 
manufacturers use higher allowable stresses for certain materials in one or another of 



Fig. 33. —Standard Diagram used in the Deriva¬ 
tion of the Machine Design Formulae. 
















72 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


the machine parts, it can only be justified on the ground of superior material. The 
assumption of superior material cannot, however, be generalized. It is to be remem¬ 
bered that, in many static strength computations, stiffness, i.e., safety against elastic 
changes of form, is more important than mere strength. A cross-head guide, for 
instance, or a crank-shaft can, in spite of sufficient strength, give rise to serious 
interruptions of service if not designed with proper regard to stiffness or to torsional 
deformation. This point is especially important in the case of sliding or rotating, 
machine parts and for surfaces supporting any important packings. 

Be on the safe side! Far-fetched refinements have no place in strength compu¬ 
tations; because the estimation of the cause and of the magnitude of the stresses in 
constructions not yet built depends generally upon assumptions which in reality are 
never quite true. At the outset even the properties of our principal materials of 
construction are subject to uncertain variation. Internal stresses and changes induced 
in them during the process of manufacture and afterward escape definite consideration 
altogether. Considering this and the occasional unavoidable accidental occurrences in 
operation, it will be clear that in our strength computations, the factor of safety should 
be considered of more importance than the factor of “accuracy,” and that, for the 
rest, the main aim should be to make the work as clear and to the point as possible. 


B. DETERMINATION OF THE PRINCIPAL DIMENSIONS 

I. Cylinder Diameter D and Stroke S on the Basis of Thermodynamic Laws 

It is apparently but a step from the fundamental thermodynamic laws discussed in 
Part II to the determination of the cylinder dimensions required for any given engine 
capacity, but just here is where our present thermodynamic knowledge fails us. It is- 
true, of course, that the theoretical pressure diagram of the gas engine is almost 
definitely defined, and the area developed so logically fixed that it is quite easy to 
determine the amount of work done either by means of it or by the aid of fundamental 
equations. But this does not serve the designer, since he has to deal not with the 
theoretical but with the actual indicated diagram. The actual diagram, however, is 
subject to so long a list of accidental variation's (as gas content, purity, and temperature 
of mixture, state of mixing and kind of combustion, location and effect of ignition, form 
of combustion chamber, amount of the heat losses, etc.) that it is next to impossible to 
draw even approximate comparisons in this case between theory and practice. If this 
is attempted we are apt to end up with a number of scattering factors of estimation 
and judgment which make it easy to compute any desired result except the correct 
one, and to which the man at the drawing board becomes accustomed only with 
difficulty, because the deductions are mostly based on the somewhat remote ground of 
strict thermodynamic theory. 

The ratio of the actually indicated work to that shown by the theoretical diagram, 
which in conformity with steam-engine practice we will call the card factor , varies 
between .4 and .8, i.e., the variation is not less than twice the lower limiting value. 
Under such conditions no reliable estimation of the true value is possible. If therefore 
it is desired to obtain easy and practically useful design formulae for the determination 


DETERMINATION OF THE PRINCIPAL DIMENSIONS 


73 


of the principal dimensions, their derivation should not be based on the theoretical 
card or upon the general fundamental equations for gases. The designer is then forced 
to use a practical method of computation in which the accuracy of the final result 
depends, as usual, first upon the judgment of the estimator, and secondly, upon the 
correct inter-relation of the various factors of experience that enter the problem. 


II. Dimensions D and S according to the Quantity of Air required for 

Combustion 

The simplest and most certain basis for the determination of the cylinder dimen¬ 
sions required for any given engine capacity is found in the amount of air necessary 
for combustion. Air is power for the engine builder, for the two are in direct 
proportion. (Doubling the weight of air drawn in allows of a doubling of the weight of 
fuel, increasing the engine capacity in the same ratio). Now the amount of air per 
stroke available for combustion depends only upon the details of construction, and 
since these are under the control of the designer in every case, the method of computing. 
the possible capacity of the engine from the amount of air used in unit time, or, 
inversely, of computing the cylinder dimensions required to generate the volume equal 
to the certain volume of air necessary for any given capacity, is to him the most 
natural. This method of computation allows of the fullest consideration of the properties 
of all the various fuels; it requires but very few assumptions, and those that must be 
made are easy of clear determination in the light of experience. 

In the following, a method of computation for the determination of the cylinder 
dimensions based upon the considerations above stated, is briefly outlined. Its practical 
applicability to gas engines of all types and sizes has been demonstrated for some 
years. The scheme incidentally has another advantage in that two values (Lh and C \), 
which are used later on, for the determination of valve and pipe dimensions, are 
determined at the outset, once for all. 

1. General Fundamental Equations. 

Let Nn = nominal brake horse-power; 

n = r.p.m.; 

D= piston diameter in feet; 

S = stroke of piston in feet; 

Vh = .785D 2 S = piston displacement in cubic feet; 

V=fj e Vh =the volume of mixture under standard conditions actually drawn in 
(29.9" Hg. and 32° F.); 

L = the practically most favorable amount of air required per cubic foot of gaseous 
or per pound of liquid fuel, in cubic feet. This is not the theoretical amount 
of air required for the fuel, but the theoretical plus such excess as seems 
best in practice; 

Lh = the actual amount of air required in cubic feet per suction stroke of the engine, 
determined from L and for the normal power N n ; 

Cs = the amount of fuel used per hour at the normal load N n , for gases in cubic 
feet, for liquids in pounds; 

C = the amount of fuel per H.P. hour (otherwise same conditions as for 

Cs); 

Ch= the amount of fuel per suction stroke (other conditions as under C«); 


74 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


H = the lower heating value of the fuel in B.T.U., for gases per cubic foot, for 
liquids per pound; 

V 

ije—~y ~ = volumetric efficiency of the suction stroke (referred to standard condi¬ 


tions) ; 

Tjw =economic efficiency = thermal efficiency at the brake = 
For four-cycle engines (per cylinder end): 


NnX 33 000X60 
778 CsH 


2545 N n 
CsH ‘ 


N n X 33 000 X 60 _ 25452V». 
77 8Ht) W Hr] w ’ 


( 1 ) 


_ 2C S __ 2X2545Nn _84.8N, t , 

h 60n 60 nHrjw nHrj w . . ^ 

CsL 2545 NnL 84.8 N n L 

Lh 30 n 30 nHrjw nHrj w ‘ . 

For two-cycle engines equations (2) and (3) should be divided by 2 , because 
there is a charging stroke every revolution. 

(a) Engines for Gas Fuel. The actual charge drawn in during one suction stroke 
must be 

V = Ch + Lh. 


This volume requires an actual piston displacement of 


V h = .785D 2 S = ^^ 1 

Tje 


84.8Nn+84.8iVttL 

nHljeljw 


84.8AT W (14-L) 


cubic feet.(4) 


nHf) w T)e 

Solving (4) for D, S, and n in turn, we have 

.< 5 > 

o 108N n (l + L) e 

o= — y rF . 0 -feet.(6) 

nHD 2 rj w T)e 

108N n (l+L) 

n= HD 2 ST)wT)e~ r - P>m . (7) 


( b ) Engines for Liquid Fuel. In these engines the fuel is introduced into the 
cylinder either as liquid or vapor. But even in the state of vapor the ratio of fuel 
volume to air volume in the mixture is much smaller than it is with the richest 
gases. For example, in the case of illuminating gas, about 5.5 volumes of air per 
volume of gas are theoretically required; for benzene vapor the ratio is 45 to 1. Even 
crude alcohol, comparatively low in heating value, theoretically occupies only about 4% 
of the volume of the mixture, and on account of the excess of air the real figure is 























DETERMINATION OF THE PRINCIPAL DIMENSIONS 


75 


between 2 and 3%. In view of this fact the volume occupied by the vapor in the fuel 
mixture may be neglected, and we may write directly 


Vh = .7S5D 2 S = — 

Tie 


84.8 N n L 

nHrjwTje' 


( 8 ) 


Solving (8) for D, S, and n, we obtain for engines using liquid fuel, 


n A llOHNnL , . 

D = \-iTa -feet;. 

*nHbrjwTje 

..... (9) 

„ 108 NnL , . 

nH D^TjuiTje 

.(10) 

108 NnL 

.(ID 

n HD 2 Sriwr)e T ‘ V ' m . 


Equations (4) to (11) do not contain a single factor that may not in every case 
be determined with accuracy. Depending completely on judgment only are the factors 
rjw and fj e , and in view of the wealth of experience available, a mistake with reference 
to these two is hardly possible. Every designer knows what thermal efficiency on the 
brake (t) w ) he may demand for the given size of engine and the given kind of fuel. 
A moment’s consideration will also show him that, for instance, for a high-speed 
X machine with automatic inlet valve and imperfect cooling, the volumetric efficiency 
(rje) must be chosen lower than for good engines of larger size. Values of based 
upon experience will be found on p. 31, Part I. For r) W , H, and L, Table 7, p. 76, 
gives average practical figures. In this table the figures in columns 3 to 8 refer to the 
nominal H.P. ( N n ) of the engine. Since this capacity N n in good constructions is from 
about 15 to 20% less than possible maximum capacity, and since for this V max a sufficient 
excess of air is required, the amount of air required for N n should never be taken into 
computation at less than 30% excess over the theoretical amount. But high-compres¬ 
sion pressures, especially in connection With the rich liquid fuels, require a still further 
dilution of the mixture, and amounts of air in excess from 50 to 6C% are common. 
For obvious reasons the amount of air used can be nearer the theoretical the more 
perfect the mixture, the lower the compression, and the more effective the cooling. A 
perfectly uniform mixture is difficult to obtain in the case of liquid fuel engines; the 
fuels used by them are very rich throughout, and, on account of the usually very 
complex charging actions, they are much more sensitive against overload than other 
machines. All of this compels the use of a greater excess of air in liquid fuel engines 
than in others, and this has been taken into account in the figures given in the table. 

The fuel consumption, C (columns 4 to 8) presupposes up-to-date construction and 
proper practical conditions of operation; on the testing floor and in acceptance tests 
better figures are very often obtained. Of course there are still several types which 
appear to be satisfied with lower economy. The consumption of ignition and heating 
apparatus has no connection with the internal operation, and has consequently not 
been taken into account in determining C. In the case of power-gas engines it is 
assumed that suction generators are used, in which case separate steam-boilers are not 
required and the generator itself furnishes the heat required for the vaporization of the 
water. The consumption of the gaseous fuels is based on 28.9" Hg. pressure and 59° F. 










Table 


76 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 



1 Including 10-15% of the full daily consumption for firing-up. The consumption of a generator banked through the night is from one-half to 

one-third of this amount. 


























































































































DETERMINATION OF THE PRINCIPAL DIMENSIONS 


77 


2. Special Equations for the Principal Fuels. From the fundamental equations (4) 
to (11) we may derive useful special equations for the main gas-engine fuels by sub¬ 
stituting in them average values for the heating value (77 ) and for the air required 
(L). Only tjhe factors r/ w and rj e then remain unknown, and these values for the 
given conditions may be taken from the tables. In the derivation of the following 
special equations (12) to (29), which apply to slow-speed stationary engines of fair 
quality, one step further has been taken in that rj e has been assumed = .85 for this type 
of engine. For this kind of machine, therefore, the designer is called upon merely to 
estimate the value of rjw for the size of the engine and to determine the fuel to be 
employed. For types differing from the above, high-speed automobile engines, for 
instance, the fundamental equations (4) to (11) should be used, on account of the smaller 
values for rj w and ij e . 

(a) Engines for Gas Fuel. —Illuminating Gas (from bituminous coal): 


77 = 565 B.T.U. per cu.ft.; L= 8.5 cu.ft. per cu.ft. 



( 12 ) 



(13) 



(14) 


Power Gas (Producer Gas): 


77 = 135 B.T.U. per cu.ft.; L = 1.30 cu.ft. per cu.ft. 



(15) 



(16) 



. . (17) 


Blast-furnace Gas: 


77 = 108 B.T.U. per cu.ft.; 7) = 1.1 cu.ft. per cu.ft. 



( 18 ) 



. (19) 



. ( 20 ) 






















78 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Examples. 1. Illuminating-gas engine, Otto-Crossley. 1 N n = 35 B.H.P., D = 13.2" (1.1 ft.). 
£ = 20.9" (1.74 ft.), n ~163 r.p.m., H = 547 B.T.U. per cu.ft., iju> = .208. Assume L =8.5 cu.ft., and 
t\e = .85 Then according to equation (5), 


D 


According to special eq. (12), 


-4 


108X35X9.5 


163 X 547 X1.74 X .208 X .85 


— = 1:14 ft. = 13,7". 


D 


■ I 2 . 

= ' 1.74X 


2.14X35 


4 X163 X.208 


= 1.126 ft. = 13.5". 


2. Old Type Producer-gas Engine, Otto-Deutz. 2 An = 150 B.H.P. in two cylinders, D = 20.5" 
(1.71 ft.), £ = 30" (2.5 ft.), n~140 r.p.m., ■q w = .l\b on the basis of coke = about .16 referred to 
producer-gas. H~ 130 B.T.U. per cu.ft., assume L = 1.25 cu.ft. per cu.ft., and ij e = .85. 

According to eq. (5), 


D 


-4 


108X75X2.25 


140X130X2.5X.16X.85 


- =1.72 ft.=20.6' 


According to special equation (15), 


D 


=v= 


2.16X75 


2.5X140X.16 


= 1.71 ft. = 20.5". 


3. Modern Suction Gas-engine, Korting Bros., JVn*=100 B.H.P., D = 20.5" (1.71ft.), £=27.6" 
(2.3 ft.), n = 140 r.p.m. 

With the general assumptions made, we will have by special eq. (15), 


D=\ 


' 2.16X100 
2.3 X140 X.24 


= 1.68 ft. = 20.2". 


4. Suction-gas Engine, Glildner. A„ = 100 B.H.P., D = 18.7" (1.56 ft.), £=27.6" (2.3 ft.) 
fi = 160 r.p.m. 

From special eq. (15), 




16X100 


2.3 X160 X.24 


= 1.53 ft. = 18.4". 


5. Blast-furnace Gas Engine, Korting Bros. 3 A M = 100 B.H.P., D = 19.7"(l 64 ft) £ = 34 5" 

(2.87 ft.), n. = 130 r.p.m., R = 112 B.T.U. per cu.ft., C = 91.7 cu.ft. per B.H.P. from which ^ = .24. 
With jj e = .85 and L = l.l cu.ft. per cu.ft., we will have, according to eq. (5), 


D 


-4 


108X100 X 2.1 


From special eq. (18), 


130X112X2.87X.24X.85 


- = 1.63 ft. = 19.6" 


\0 07 


.47X100 


2.87X130X.24 


— = 1.66 ft. = 19.9" 


Since eq. (18) assumes A-108 B.T.U. in its derivation, the value for D is in this case 
naturally somewhat greater than by eq. (5). 


1 Schottler, Die Gasmaschine, 3d ed., p. 167. 
* Schottler, Die Gasmaschine, 3d ed., p. 172. 
8 Stahl und Eisen, 1900, p. 413- 


















DETERMINATION OF THE PRINCIPAL DIMENSIONS 


79 


ft.), 


6. Blast-furnace Gas Engine, Cockerill. 1 # n = 600 B.H.P., D=51.2" (4.26.ft.), 5=55.0" (4.59 
n = 95 r.p.m., # = 108 B.T.U. per cu.ft., tj w = .19, ij e and L assumed as before. Then from (5), 


D-4 

According to special eq. (18), 


108X600X2,1 
95 X108 X 4.59 X.19 X.85 


= 4.24 ft. = 51.0". 


D 


-4 


2.47X600 
4.59 X 95 X.19 


= 4.25 ft. = 51.1" 


(6) Engines for Liquid Fuel.— Kerosene: 

# = 18 900 B.T.U. per lb.; L = 300 cu.ft. per lb. 

-\ /2.02iVn f , /oi\ 

D= ^h®^ feet; . (21) 

„ 2.02 N n t , /oox 

5= ^r feet; . (22) 

2.02 Nn 

” = D^ r - p - m . (23> 

Gasoline: 

#=19 000 B.T.U. per lb.; L= 280 cu.ft. per lb. 

D = feet;.(24) 

TL&7]u) 

S = 1 J^ n ieet; .(25) 

nD 2 Tj w 

1.87 Nn /oa\ 

n= ws^ T ^ m . (26) 

Crude Alcohol, about 90% by volume: 

# = 10 250 B.T.U. per lb.; L = 160 cu.ft. per lb. 

. (27) 

.. • • • {28) 

.< 29 > 


1 Stahl und Eisen, 1900, p. 721. 


























80 


DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Examples. 1. Kerosene Engine, Grob & Co. 1 N n = 8 B.H.P., D — 9.05" (.75o ft.), 5 — 9.05" 
(.755 ft.), n = 266 av., H = 19 850 B.T.U. per lb., j? w = .136. Assume L = 300 cu.ft. per lb., and on 
account of automatic inlet valve J? e = .83. Then according to eq. (9), 


D 


- 4 . 


108 X 8 X 300 


and from special eq. (21), 




266 X19 850 X .755 X .136 X .83 

= .77 ft. = 9.25". 


- = .76 ft.=9.0". 


2.02X8 


266 X.755 X.136 


2. Alcohol Engine, Altman & Co. 2 D = 11.0" (.92 ft.), 5 = 15.8" (1.31 ft.), n = 200". With 
alcohol alone N n was 18 B.H.P. With a benzol-alcohol mixture r)w varied between .13 and .17. 
Assuming the same efficiency of heat transfer gives an average — for alcohol alone. Then, 
with L = 160 cu.ft. per lb., eq. (9) gives 


d=4 


108X18X160 


200 X 10 250 X 1.31 X .15 X .85 


- = .954 ft. = 11.45". 


For special eq. (27), 


/— 1.98X18 Q5 ft> = 1L4 o". 

v 200X1.31 X.15 


The close agreement of the computed with the actual dimensions shows the 
general applicability of the method above developed. In order to extend it to 2-cycle 
machines it is only necessary to introduce into the formula a factor representing the 
ratio of capacities. This type of machine works throughout with special charging pumps 
whose discharge capacity of course directly affects engine capacity. With pumps of 
sufficient size a means is given of considerably increasing the weight of air per suction 
stroke for the combustion process by merely raising the transfer pressure. The effective 
power of the engine, however, does not increase in the same ratio because the increase 
in the air supply entails a higher exhaust pressure and increased work in the pumps, 
which means a decrease in the mechanical efficiency of the engine. Both of these 
effects can be taken into account by suitably changing the value of t) w . 

The special equations give an insight into the ideal ratios of specific engine 
capacity for the various kinds of engines, assuming t) w the same for all types. To 
judge of the real ratio of capacity, the actual value of t] w must be considered. Thus, 
for example, the ideal capacity ratio between an illuminating-gas and a producer-gas 
engine, assuming rj w equal, would be as 2.16:2.14, from eqs. (12) and (15). The difference 
is very small. In practice, however, the former may be expected to give a value of 
rjw of about .27, while the latter only gives .24, so that in reality the producer-gas 

.27-.24 

engine would have to be larger than the illuminating-gas engine by about -— = 10 

to 12%. This does not mean that it is not possible to obtain the same B.H.P. from 
any given illuminating-gas engine when operated with producer-gas, if only the engine is 
of ample size for the power and the combustion is not too imperfect, so that the values 
of i) w are not too far apart. It is a fact, however, that these conditions are found 
only in a few excellently designed and well constructed machines. 


1 Z. d. V. D. I., 1895, p. 616. 


2 Ihering, Die Gasmaschine, 2 ed., p. 301. 











DETERMINATION OF THE PRINCIPAL DIMENSIONS 


81 


III. Dimensions D and S according to Practical Capacity Coefficients 

The following practical capacity coefficients are all based on experience with 
machines actually constructed. Their use therefore presupposes similar conditions. For 
special cases, or for those which require strict treatment, these empirical coefficients can 
be relied upon to only a small extent. In such cases it is better to use the method 
outlined under II. 

Table 8 gives the most common capacity coefficients. The figures refer to the 
nominal brake horse-power, N n , of 4-cycle machines. For maximum load, therefore, 
they should be increased in corresponding proportion. For 2-cycle engines the specific 
capacity L 0 of 4-cycle machines (that is, the work done per second per cu.ft. of dis¬ 
placement) may be multiplied by from 1.75 to 1.90. 1 The specific piston displacement 
V 0 (that is, the cubic feet piston displacement required per second to develop one 
horse-power) is thereby decreased in the inverse ratio. 


Table 8 

CAPACITY COEFFICIENTS FOR 4-CYCLE ENGINES 


Assume Mechanical Efficiency 

^=.80. 

M.E.P. 

Pi- 

Pounds per 
Square Inch. 

Pn~Pii)m- 

Pounds per 
Square Inch. 

144p„ 
Lo - r -. 

Foot-pounds. 

T . 550 

Cubic Feet. 

K* = .0017p n 

Relative 
Capacity 
Referred to 
Illuminating- 
gas Engine 
= 1.00. 

Illuminating-gas engine. 

78.0 

62.5 

2250 

.245 

.106 

1.00 

Producer-gas engine. 

60.5 

48.4 

1740 

.316 

.082 

.77 

Blast-furnace gas engine. 

57.0 

45.6 

1640 

.335 

.078 

.73 

Gasolene engine. 

74.5 

59.6 

2150 

.256 

.102 

.95 

Kerosene engine. 

54.5 

43.6 

1570 

.350 

.074 

.70 

Diesel oil engine. 

99.5 

79.6 

2870 

.192 

.136 

1.27 

Alcohol engine. 

54.5 

1 

43.6 

1570 

.350 

.074 

.70 


* See page 83. 


By substituting the above capacity coefficients in the formulie following (eqs. 
30 to 37), the principal dimensions may be determined. The notation used in the 
derivation of these formulae is the same as that given on p. 73. 

1. Indicated mean effective pressure pi in lbs. per sq.in. 


Nominal B.H.P., N n = 


U4pir)mX.785D 2 Sn 

33 000 -2 


PiT)mD 2 Sn 

585 


(30) 


1 Translator’s Note. It is evident that this statement can apply to medium and large sized 2-cycle 
machines with separate pumps only. It is a well known fact that the power of small high-speed 2-cycle 
engines with crank case compression is often no greater than that of 4-cycle machines of equal 
cylinder diameter. This is due mainly to low volumetric efficiency of the small 2-cycle cylinder. 
























82 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Or, if we let F = area of piston in sq.ft., and c=-^=piston speed in feet per second, 
we may also write 

Nominal B.H.P., N t = J^^^- = .066p^Fc = .066piTjmU„ . . . (30a) 

in which U S = F C = ‘ 785 ^ S - cu.ft. = piston displacement per second. 

oU 

From eq. (9), p. 8. we know that the piston displacement Vh required per stroke 
for Ni indicated horse-power at n r.p.m. is 


V h = 


33 000Ai 458Ai 


144p5 


pm 


cu.ft. 


For one horse-power , therefore, the required displacement per stroke is 


-wj 458 
V h — — cu.ft., 

Pin 


(31) 


and for any given cylinder dimensions the indicated horse-power is 

. 


(31a) 


It is evident that the factor expresses the I.H.P. generated by one cubic foot of 

458 

458 

piston displacement in 4-cycle engines, and conversely measures the cubic feet of 

piston displacement required for every I.H.P. These two reciprocal factors become 

equal to unity, that is, one cubic foot of piston displacement will generate one I.H.P., 

when pin = 458. In the average machine with pi = say 75 lbs. per sq.in. and n~180 

r.p.m., pm will equal about 13 500, which means that each cubic foot of piston displace¬ 
ment in a 4-cycle machine will furnish about 29 I.H.P. 1 

If the engine operates on the hit-and-miss system of regulation so that there are z 
explosions per minute, the I.H.P. of any partial load will be 


Ai = 


144piX.785D 2 Nz 
33 000 


-.0034 piD 2 Sz. . (32) 


Assuming that the indicated mean effective pressure of the suction and exhaust 
strokes combined is p w lbs per sq.in., the work of charging and discharging the 
cylinder will be 


A» = 


144pu,X.785D 2 s(|-z) 


33 000 


= .0034 p w D‘ 


113 

1 For single cylinder steam engines eq. (31) has the form Vh = ——. For the average case pt — 30 to 

J>in 

40 lbs. per sq.in. When it is further considered that the economical number of turns for the steam 
engine is about one-third less that for the gas engine, the ratio of the specific piston displacement between 
gas and steam power will be about as 1.0 to .75, which is not as unfavorable to the gas engine as is 
generally assumed. 




(33) 












DETERMINATION OF THE PRINCIPAL DIMENSIONS 


83 


Consequently the net I.H.P. of the engine will be 

Ni' = Ni — N w = .0034 D 2 ,S’ [zp ( - (j - *) P»].(32a) 

The indicated I.H.P. gives a higher mechanical and a lower thermal efficiency than 
the gross I.H.P., since Ni>N t '. In regard to this see page 8 f 

2. Useful mean effective pressure p u lbs. per sq.in. for N n B.H.P. 


or 


. 

Nn = = .066 p n Fc = .066 p n V 8 . 


550X4 


(34) 

(34a) 


Conversely, therefore, 


N n Nn 

Pn =:omFc = mv , lbs - per s< >’ in - 


3. Specific capacity, L 0 , in ft.-lbs. per cu.ft. of piston displacement per second. 


Nominal B.H.P., N n =- .000048 L 0 D*Sn - ^ .(35) 


4. Specific piston displacement, V 0) in cu.ft. per second for 1 B.H.P. 

. 1DUT) ,, .785D 2 Sn Fc D 2 Sn 

Nominal B.H.P., — . . 


from which 

5. Capacity constant K: 


v - Fc = 1)2Sn 

0 Nn 38.5 Nn 

N n = KD 2 Sn B.H.P. 


(36) 


(37) 


in which the constant K = qqq = -0017pn from the general equation 

_ 144p n X.785D 2 Sn 
n 33 000-2 


Eq. (37), on account of its great simplicity, is especially well adapted for rapid 
computation or for comparison of capacities. 

A useful aid for the determination of D and S is given in the diagram, Fig. 34. 
The curves are constructed for a mean p n = 65, or p t = about 77.5 lbs. per sq.in. 
and for 200 r.p.m. The best way to use this diagram is to determine first not the 
nominal but the maximum capacity which the engine is to have (since 77.5 lbs. per sq.in. 
is near the upper limit for average construction) and then to find proper values of D 

and p — from the curves. If instead of 200 r.p.m., the figure assumed in the 

diagram, any other number of turns is desired, the maximum capacity mus„ be adjusted 
to this figure. 


Example. A 100 B.H.P. engine is to have an excess capacity of 20%, at 150 r.p.m. This 

means a maximum B.H.P. of 120, but this must be changed to 120^^=160 B.H.P. to get it to 

S 150 

a 200 r.p.m. basis. Assuming />=—= 1.6, the upper set of curves under B.H.P. = 160 gives 

D = 19.9" = say 20". The data for the 100 B.H.P. engine will then be D = 20", <o=1.6 £=1.6X 
20 = 32", r.p.m. = 150. 
















■f- y3 pi = 77.5 lbs. per sq.in. 























































































































































































































































































GENERAL ENGINE PARTS 


85 


C. GENERAL ENGINE PARTS 

By general parts are meant those used for gas as well as for oil engines; according to 
the kind of fuel employed these are supplemented by special parts, which are taken up in 
the next chapter. Finally, every engine installation requires certain auxiliaries which, 
although they are not a part of the machine proper, are generally built and furnished 
by the engine manufacturer, and are for that reason considered in Part E, p. 268. 

The designs shown in the following construction plates are for the main part 
taken from modern and tried types of engine, but they should not on that account be 
taken as models without exception. Special circumstances sometimes require and 
sanction constructive means, which are inferior or remain of doubtful value as far as 
general application is concerned. As a caution it should be stated that some, although 
comparatively few, of the types and general details shown are coyered by patent, and 
are therefore not open to general use. It was not possible to obtain reliable informa¬ 
tion regarding the field so protected, useful as this would have been to the designer, 
because definite statements on this point could be obtained from only a few of the 
manufacturers. 


I. Beds and Frames 

Material. Mostly cast iron, for light automobile engines also aluminum alloys, and 
now and then cast steel or soft steel; for crank bearings of the smaller sizes nearly 
always brass or bronze (phosphor-bronze), for other sizes cast iron lined with white metal. 

Allowable Fibre stress is comparatively low, in consideration of the importance 
of this machine part and the generally uncertain estimation of the stresses occurring. 

In the case of stationary engines this allowable fiber stress for cast iron of the best 

grade should not exceed 3500 lbs. per sq.in. in bending, as computed from the total 
explosion pressure, P z . 

The body of vertical engines is usually called the frame, that of horizontal 
machines usually the bed, although the use of these terms is not at all strictly defined. 
For both of these in the course of time fairly similar fundamental forms have been de¬ 
veloped, which are pretty closely adhered to unless necessity demands other constructions. 

In every machine part under stress it is important that the stress be taken up 
as far as possible in a central or axial direction. To take up the maximum total 
explosive stress P z in this way is an especially important requirement in the 
design of beds and frames. In the case of frames this can be fulfilled easily and 
by simple means, in the case of engine beds, however, only by employing designs 

as yet little used. (See Figs. 83-86, p. 106.) The types of horizontal engine beds 

generally employed are all subject to considerable bending stresses. In every case the 
moment arms of the bending stresses, where these cannot be avoided, must be made 
as short as possible if a safe construction is to be obtained with the least expenditure 
of material. It should be considered that, with maximum total explosion pressures 
ranging from 300 to 400 tons, and such are found in the large engine, every of an 
inch off or on the length of a moment arm, has a considerable influence on the size of 
the bending stress. To relieve the bed or frame through connections with pedestals or 
foundations is of doubtful utility, since such connections are always more or less 
yielding, and often accompanied by other and unknown stresses. Relief in that way 
should never be sought and should not in any case be taken into account in strength 


86 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


computations. The machine bodies themselves, therefore, should under all circumstances 
be able to take up all stresses occurring during operation. Twisting of frame during 
manufacture or erection can be quite successfully prevented by sufficiently large base 
surfaces, structural stiffness of the casting and proper location of the foundation bolts 
or anchors (close to vertical walls and near the points of support). 

The unit pressure or load on foundations due to weight of machine and whatever 
inertia forces there may be, should not exceed the following figures for the various 
materials employed: 


Lbs. per sq.in. 


Granite blocks. 110-140 

Sandstone. 70- 80 

Hard-pressed brick in cement.... 55- 70 


Lbs. per sq.in. 


Brick and concrete. 42-55 

Hardwood timber. 55-70 

Pine and hemlock. 30-40 


It is well to keep the unit load from one fourth to one third less than these 
figures under the crank bearings of horizontal machines. This gives a means of 
computing the necessary area of the supporting surfaces. 

For dimensions of foundations, see p. 258. 

1. Frames for Vertical Engines, Crank-shaft above Cylinder (Figs. 35-43). This 
type of frame is to-day used only for small powers, up to 6 H.P., and for speeds not 
to exceed 250 r.p.m. It possesses the advantage of a fly-wheel running entirely free, 
and needs little or no foundation. But for larger units these frames have too little 
stability, owing to the high location of the gravity center, and there are other 
disadvantages in static respects. 

Designs. 



Figs. 35-38. —Frame for a 

2-cycle Giildner Engine. 

(The upper end of the 
cylinder serves as a charge 
pump, which, by means of a 
port a, communicates with 
an annular space a used 
as a receiver- The charging 
and discharging actions of 
this pump are controlled by 
a slide valve on the top of 
the upper cylinder cover, 
which valve is operated by 
the connecting rod passing 
through an oscillating ball 
joint in the valve itself. 
Space b is the exhaust 
chamber, c the jacket space. 
After the exhaust gases have 
escaped into b the inlet 
ports from a are opened.) 












































































































GENERAL ENGINE PARTS 


87 



Figs. 39-43.—Frame for a Kjelsberg Kerosene Engine, built by L. Nobel, Petersburg. 

(The web projection on the cylinder head is aised to prevent the piston from dropping down too far, 
in which case the piston rings may catch behind the counter-bore.) 

Constructive Details. Length of stroke, S, generally from 1.75 to 2.0 times the 
cylinder diameter D, for larger sizes down to 1.5 D; height to center of main bearings 
from 4 to 5 S, and length of side of the generally square base from 4.5 to 5 D. 
Prismatic frames of this type are now used but very little, although the cost of the 
patterns is less and the weight of the frame 10% less than in the case of cylindrical 
frame bodies of equal capacity. Generally cylinder and frame are cast in one. Separate 
cylinder barrels cause some difficulty on account of the lateral connection with the ports, 
and from considerations of strength require a thickening of the walls of the frame. 

Strength computation results in thickness of cylinder and jacket wall too thin for 
practical work; on this account with special reference to foundry practice, the dimen¬ 
sions may be made as follows for one-piece frames: 

For 7) up to 5.0", cylinder wall (s) = .50-.60"; frame walls (s') = .40" 

“ 7.0", * “ “ (s) == .60-.70"; “ “ (s') =.45" 

“ 8.0", “ “ (s) = .70-. 80"; “ “ (s') = .50" 

“ 9.0", “ “ (s) = .80-. 90"; “ “ (s')=. 55" 
















































































88 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


With the above dimensions the tensile stress in the cylinder wall, as computed 
from the explosion pressure, and without reference to the influence of the frame, will 
be only about 1400 lbs. per sq.in. If separate cylinder barrels are used, the thickness s 
of the cylinder wall may be reduced to the figure found by computation [eq. (11),. 
p. 124]. For the design of the bottom end of the cylinder see p. 134. 

In the case of the larger frames, Figs. 39-43, the transition section between the 
bearing support and the jacket wall must be specially checked; because at this point 
p . 

the force is not taken up in an axial direction, but acts through a moment 

arm, a. It may be resolved into the tensile stress, P n , and the shearing stress, P s 
(see triangle of forces in Fig. 40). The stress P s may be neglected and the computa¬ 
tion confined to the determination of the combined stress resulting from P n (tension) 

p 

and -y-a (bending). Assuming the area of the dangerous section x-y , Fig. 40, to bo 

/ sq.in., and the section modulus equal to IF, we will have the max'imum combined 
stress 


a 


Pn 

f 


Pz 

~2 a 

H—jy lbs. per sq.in. 


( 1 ) 


t The value of P n is best found graphically. The distance a from the center line of 
the bearing to the neutral axis of the dangerous section, x—y, is determined most 
easily by balancing a stiff paper model of the section concerned. See p. 90. 

Since the line of action of the fly-wheel weight G is opposite to that of the 

piston pressure P z , the opposite bearing will be relieved of its load — by a certain 

A 

amount. But the weight G referred to the center line of this bearing acts as a load 

as soon as — <G, which condition takes place during the expansion stroke. And since 

this bearing is thus loaded the greater part of the time, the fly-wheel is to be kept as 
light as possible, and seated as closely as possible to the adjacent bearing. For 
details relating to the determination of bearing pressures, see under “Crank Shafts ” 
p. 164. 

The width of the jacket space should in no place be less than 1$". The external 
shape of the frame usually gives a water space considerably wider than this, which is 
favorable to good casting. For the removal of the core sand suitable cleaning holes 
should be provided in the base. These are afterward closed watertight by fine 
threaded plugs. The closing of these core sand openings should not be made a part 
of the duty of the cylinder head packing. The curves at the base between jacket 
wall and base plate are parabolas, the construction of which is shown in Fig. 42. The 
usual method is to draw this curve once and then to use circular curves of approxi¬ 
mately the same form. 

The caps and bolts of the bearings must be designed for the maximum explosion 
pressure. The load due to the weight of the fly-wheel, added to the load on the bearing 
next to this wheel, can usually be neglected, because general considerations of safety 
and the fact that the direction of the belt pull is hardly ever definitely known, make it 

imperative to keep the stresses down to a comparatively low figure. Let P = — lbs. be 

half of the explosion pressure, also let the distance from center to center of studs or 



GENERAL ENGINE PARTS 


89 


bolts be a in. and the section modulus of the dangerous section of the bearing cap be W , 
4kbW 

then from P = ——, in which fc& = allowable bending stress, we shall have the maximum 
bending stress 

oh = lbs. per sq.in.(2) 

If the dangerous section is in the shape of a geometrical figure, a rectangle for instance, 
as is usual in small machines, the section modulus W may be taken directly from a 
table. In other cases it is necessary to first determine the moment of inertia I of the 
cross-section, which is best done by dividing the given section into simple figures, and 
for each of these determine I referred to the gravity axis of the section. The sum of 
the partial moments so found is equal to the moment I of the total section. The 
following example makes this clear. 


Example. The dangerous section of one of the bearing caps of the frame shown in Figs. 
39-43 has the dimensions given in Fig. 44. 1 Since the cylinder diameter is 6.9", the total 
explosion pressure Pz for a unit pressure of p 2 = 255 lbs. (for which this engine is constructed), 

will be P« = 255X.785X6.9 2 * = 9500 lbs. Of this load ^^ = 4750 lbs. = P will fall to each bearing. 


With the given distance a between bolts equal to 5.1", the greatest bending moment in the 
cap is 

Mh =~X^ =—^^=6100 in.-lbs. 2 
I 1 4 

— V 



For the determination of the moment of inertia the cross-section is Fig. 44. —Dimensions 
divided into three parts, I, II, and III, Fig. 44. Minor variations and in Inches, 

approximations in outline, such as made in the figure, are of no im¬ 
portance. The distance of the gravity axis from any given reference line, such as the upper 
horizontal line in the figure, may be computed as follows: 


2.36 XI.81 2 1.97X.63 2 
2 2 

p = _:__ 78" 

1 (2.36 X 1.82) + (1.97X.63) ' ’ 

hence e 2 = 1.81 — .78 = 1.03". 

For sections more complex than this the position of the gravity*axis is best located by 
cutting a proportional model out of stiff paper and balancing this on a sharp straight edge. 

The several moments of inerria of the section of Fig. 44, referred to the gravity axis will be 


For area 


A QQy «Q3 

= jPJL + (4.33X.63X.43 2 ) =.593 bi-quad. in. (or in. 4 ). 
12 


1 All dimensions given in examples will be in inches, since nearly all of our strength computations 
are based on this unit. 

2 All results are given in round numbers, due to the fact that the slide rule is used. The approxi¬ 

mations have no effect on the final results. 

















90 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


1 1O w 1 103 

For area II, In = ' * ‘ + (1.18 X 1.18 X .47 2 ) = .466 bi-quad. in. 

1 18X1 18 s 

For area III, / m = --— ^ ' — + (1.18X1.18X.47 2 ) = .466 bi-quad. in. 
So that the total moment of inertia for the section will be 


7=/i + /ii+/iii = .593 + .466 + .466 = 1.525 bi-quad. ins. 


The section modulus therefore is 



1.525 

.78 


= 1.96, 


and the greatest bending stress (in the top side of the section) will therefore be 


_m_6ioo 

° b ~ W ~ 1.96 


= 3100 lbs. per sq.in. 


The 1" cap screws each have a load of = 2375 lbs., which, according to column 6, of 
table (p. 307), is safe. 

The foundation bolts of this type of vertical frame may have the following dimensions: 


For D= 5.0" 7.0" 

Diameter of bolt= $" £" 


8.0" 9.0" 

£" 1 " 


2. Standard Frames for Vertical Engines, Crank Shaft below Cylinder. The 

so-called box frames, Figs. 45-49, are mostly used for small engines only, although 
some American firms, as Westinghouse for instance, also employ them for large 
capacities. On account of the low position of the crank shaft this type is well adapted 
to high speeds of rotation and for direct connection with the power consumer. The 
facility with which the working parts can be tightly enclosed in such a frame make it 
especially suitable for use in the open or in dusty rooms. This fact has led to the 
development of special crank cases for automobile engines (see the various constructions 
shown in Part III). Such enclosed frames in themselves insure sufficient stiffness of 
construction, if the walls are made thick enough. In special cases, however, sufficient 
webbing for the side walls must be provided, especially if for purposes of lightness 
the crank case is made of the commonly used aluminum alloy instead of cast 
lion. 




GENERAL ENGINE PARTS 


91 


Design of Box Frames. 


Figs. 45-47.—Frame with Pedestal 

for a 2 H.P. Engine (Z) = 4.75", 

8"). 

(The hit-and-miss governor and 
valve operating arrangement shown 
in Fig. 2G0, p. 206, is bolted to the 
left side of the frame near the top. 
The use of neither 2 to 1 gear nor 
lay shaft is necessary with this 
construction. Some designers use 
the pedestal as fuel tank or even 
as exhaust muffler. The latter 
practice is bad because of the con¬ 
sequent heating of the crank bear¬ 
ings. It is better to use the space 
available in the pedestal merely as 
a receptacle for tools or auxiliary 
apparatus. 



Figs. 48-49.—Frame for a Small 

Giiklner Engine, One-piece con¬ 
struction {D — 6", <S = 9"). 

(If the molding of this frame 
causes difficulty, a separate cylin¬ 
der liner may be used, casting only 
the jacket wall in one piece with 
the rest of the frame. Hand-hole 
a is used to disconnect the wrist- 
pin bearing in case the piston must 
be removed, leaving the crank-pin 
bearing undisturbed. This is of 
importance when, on account of 
lack of head room, it is not easy 
to take out both piston and rod 
together. For the valve gear be¬ 
longing to this model see Figs. 261 
.and 262, p. 206. 



- ; -->t<— zo3 —>k- /&&e 









































































































































































































92 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Constructive Details of box frames according to Figs. 45-49: 

Stroke ratio ^ = 1 to 1.75; distance from center of shaft to base from .7 to 1.0 S; 

distance center of shaft to floor from 16 to 24" depending upon size of machine and 
diameter of fly-wheel. Total height of frame above base from 5.5 to 6.5 S. Length of 
side of square base (common construction) from 3.5 to 4.5 D. The lateral openings in 
the base casting should be of ample size to allow of easy dismounting or assembling 
cf crank shaft and gear. 

The thickness of the walls of the base casting is in the smaller machines of this 
type usually prescribed by practical considerations. Based on experience, the following 
proportions may be used: 

For D = 5" 7" 9" 11 " 13" 15" 

5 = JL" 1" 11” 1” 13" 

6 5 Iff 8 16 4 it 



A check computation is required only to determine the 
tensile stress in the smallest cross-section of the frame (and 
of the jacket wall, if the construction is in one piece) and 
the bending stresses in the cross-section below the main bearing. 
This latter is especially necessary in case the construction is like Fig. 50, with the 
main bearing in a separate base plate. 


—r~ 

Fig. 50. 


Examples. 1 . With reference to Fig. 50, let P = 21600 lbs., a = 37.5", 
so that the greatest bending moment in the dangerous section Z-Z is 

,, 21600X37.5 

M b = --- =202000 in.-lbs. 


n 


Q)* 


"I? 


Irk 




-7s>y 


rit 

| -7S, 


J 

n 


M ith reference to the neutral axis of the conventional cross-section, tv_, t,. 

iff. 51. the moment of inertia of the Homra m 11 CJ csoof irvn ttti 11 Le * 1DH6D810DS 

in Inches. 


Fig. 51, the moment of inertia of the dangerous section will be 

v T . 7.9 X 1.38 3 

For area I, L=- 


12 


+ (7.9X1.38X5.55 J )=336.2 in. 4 ; 


79X11 6 3 

For area II, ' +(.79X11.6X.57 2 ) = 105.2 in. 4 ; 


For area III, 7 m = same as In = 105.2 in. 4 ; 


v nr T 2.36X1.18 3 

For area IV, 7 IV =- + (2 - 36xl - 18x7 - 0J ) = 136 -8 in. 4 ; 


3 15x1 18 3 

For area V, I v =~—^— + (3.15XI.18X7.0 2 ) =182.4. 


Therefore, 


7=2 (I n ) = 336.2 +105.2 +105.2 +136.8 +182.4 = 865.8 in . 4 


Section modulus W = - = = j 14 5 

e 1 7.57 























GENERAL ENGINE PARTS 


93 


Consequently the maximum bending stress in the section Z-Z is 


202000 

<76 = --—— =1775 lbs. per sq.m. 


2 A vertical engine, cylinder diameter 11.8", has the dimensions given 
in Fig. 52 for the cross-section of the frame under the main bearings. The 
distance between the base bolts, see Fig. 50, is 49.3". The normal explosion 
pressure, according to p. 71, will then be 

Pz = 280 D* lbs. = 280 X11.8 2 = 39 000 lbs., 
from which the load P on each bearing will be 19500 lbs. 



Fig. 52.—Dimensions 
in Inches. 


The greatest bending moment in the base plate below the main bearing is 


„ 19500X49.3 

Mb =--- 


= 240 000 in.-lbs. 


Moments of inertia of the partial areas: 

6.9 X 1.18 s 


L = 


In = 


12 

.69 X 9.7 s 
12 


+ (6.9X1.18 X5.32 2 ) =230.9 in. 4 


= 52.5 in. 4 


/in = same as for In 
8.3 X.98 s 


= 52.5 in. 4 


Jiv = 


12 


+ (8.3X.98X5.4 2 ) =238.6 in. 4 

S(/„) =574.5 in. 4 


574 5 

Hence section modulus is-=97.5, and the maximum bending stress in the dangerous section is 

5.93 

m .“ =2460 lbs 

97.5 


The compression stress in the outer fibers above the neutral axis is neglected 
because the compressive strength of cast iron is much greater than its tensile strength. 
For further computations on base plates, see p. 97. 

Sizes of foundation bolts should not be smaller than those given on p. 90 for 
frames with crank shafts above cylinders. For information regarding foundations see 
p. 257. 

For medium sized vertical engines the so-called A-frame, borrowed from steam- 
ongine practice, and shown in Figs. 53-55 and 60-62, has found extended application. 
Owing to the better disposition of material, this type of frame is considerably lighter 
than the box frame above discussed. Besides this the shaft and the open end of the 
cylinder are kept free. The substitution of separate columns (compare Figs. 56-59) for 
the front leg of the A-frame cannot be recommended for high-pressure engines, because 
this construction is less rigid and costs more. 













94 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


v 4* 


Design of A-frames. 





























































































































































































GENERAL ENGINE PARTS 


95 



Figs. 60-62.— Frame with Cross-head Guides, 10 H.P. Diesel Engine of French make. 


Constructive Details. Total height of machine with A-frame and cross-heads is 
from 8-9 S, without cross-heads from 5.5-b.5 S. Other frame dimensions approximately 
as given for the previous type of frame. Batter of the outside walls of the columns 

of the frame in the ratio of 1:4 or 1:3, depending upon the ratio That is, in 

Fig. 55, angle a- should be in the neighborhood of 70°. A careful check computation 
is necessary for sections x-x (tension), y-y (tension and bending) and for the middle 
section z-z (bending) of the crank shaft-bearing, Fig. 62. 

In the minimum horizontal section x-x, of area / sq.in., we have the tensile stress 


<7 = -y lbs. per sq.in. 


(3) 


Examples. 1. Let the cross-section y-y of a frame similar to Fig. 62 have the dimensions 
of Fig 63. The cylinder diameter is 10.8", so that normal P 2 = 280 D 1 2 = 280X10.8 -32800 lbs., 
of which one half, P = 16400 lbs., falls to each column. At the base of the column P is 
resolved into the two forces P n and P b which put the section y-y under tension and bending. 


1 Usual assumption which gives excess safety, 

than this on account of the rigid connection of 
computation satisfactory data is lacking. 


In fact the bending stresses will be much smaller 
the column with the base plate. But for closer 















































































































96 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


The values of these forces are best determined graphically from the force parallelogram, Fig. 62. 
Suppose P n = 14250 lbs., and Pb = 5050 lbs., and let the moment arm 1 = 23.6". Then for the 
section y-y of area /= 34 sq.in. the tensile stress due to P n will be 





14 250 ..... 

a = ——— = 420 lbs. per sq.in. 
34 


7S3 


The moment of inertia of the section referred to the neutral axis, Fig. 63, is 


/ = ~(10.2X7.88 3 ) —(7.84X5.92 3 ) =280 in. 4 

lw 


Fig. 63.—Di¬ 
mensions in 
Inches. 


The section modulus therefore is 


W = 


280 

3dJ4 


70.8. 


Hence maximum bending stress 


a 


b 


23.6X5050 

70.8 


= 1680 lbs. per sq.in. 


Total maximum stress in section y-y is therefore 


o + o b = 420+1680 = 2100 lbs. per sq.in. 1 


As a matter of fact the stress in the section y-y is from 3-5% smaller 
the weight of cylinder and frame. 

2. For a similar A-frame, the cross-section y-y of which is shown in 
Fig. 64, let R? = 26400 lbs., so that P= 13200 lbs., and from this by 
means of the force parallelogram P„ = 12700 1bs. and Pb = 3960 lbs. The 
moment arm 1 = 21.6", and the area of cross-section /=26.7 sq.in. 


Tensile stress et= 


Pn 

/ 


12700 

26.7 


= 475 lbs. per sq.in. 


than this owing to 



Fig. 64. —Dimen¬ 
sions in Inches. 


1 Since for cast iron the allowable bending stress (Kb) is greater than the allowable tensile stress 
( K z ), we should for this metal strictly satisfy the requirement that 

Po+Oi<Kb or c+^<j£, 

0 = *' 


For the most common shapes of cross-section, = ~ = 1.2 to 1.5. With the latter value the above 
example would have given for the maximum total bending stress in the section the value 


<r, = (1.5X420) + 1680 = 2310 lbs. per sq.in. 


The difference is evidently insignificant and remains so as long as a is comparatively small as compared 
with o b . This holds good for all the examples following (except Example .3, p. 97) for wdiich reason 
Kb has been assumed equal to K z throughout, i.e., /?= 1.0. The inaccuracy introduced thereby increases 
the factor of safety as long as the relation a + (J b < K z is satisfied. 


















GENERAL ENGINE PARTS 


97 


Moment of inertia of section referred to neutral axis 
, 5.9 XI.56 3 , 

A-—-+ (5.9 X1.56 X 2.48 2 ) =58.57 

1 37x6 3 s 

Iu and h ii, each = — '' +G.37X6.3X1.45 2 ) =46.80 

l 5 o J 

Modulus of section W = ———=47.0. 


Bending moment ilf6 = 21.6X3960 = 85500 in.-lbs. 
so that the maximum bending stress, in the external fibers of the column 

85 500 


Total 7=152.17 in. 4 


a b =- 


47 


= 1830 lbs. per sq.in. 


Hence total stress is a+o b =475 + 1830=2305 lbs. per sq.in. 

In determining the stresses in the section z-z of the base plate, which section is 
.usually under the greatest stress, the stiffening due to the foundation is neglected 
because its effect is uncertain, and counterbalanced by the weight of fly wheel and 
crank shaft, by unknown accidental stresses and by the drawing up of the foundation 
bolts. For the same reasons any decreasing effect that the dead weight of frame and 
cylinder may have on the value of P z is also neglected. If the center line of the 
cylinder bisects the distance between the middle planes of the main bearings, the load 

p 

on each bearing will be simply P =—; in other cases P, referred to the middle of 
the bearing, can be easily determined. 


Example. 3. The cross-section z-z, through the support 
of a main bearing having ring oilers, Fig. 65, is to take care 
of a load of P = 55000 lbs. Let the distance between the 
column bolts on each side of the bearing be a = 55.2", then the 
maximum bending moment will be 

Mb = I Pa = —°°-* 55 ~ = 762 000 in.-lbs. 


The total moment of inertia 7 is equal to S(7»), Fig. 65. 
7 = 7i + 7 n + 7 m+7iv+7v + 7vi + I\n", 


hence 



Fig. 65.—Dimensions in Inches. 


/- p^‘ +1 3.80X1.57X7.85=] + p^’ +3 . 15 X1.77X6.6>] 


+ 


+ 


p3.64X.98 3 +1364x 98x2 ,26 2 j d-^- ' 78 * 1 - + .78Xl5X.39 2 j 
j~ .98X9.65 3 +>98x9t65X 3 >06 »j + +15.2 X .98 X8.35 2 j 


= 1334.4 + 243.6 + 69.1 + (2 X 220.8) +161.7 +1041.2 = 3291.6 in.‘ 



























98 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


I 3291.6 

Therefore the modulus is W =-•= -' =371.0, 

€i o.o5 

and the maximum bending stress in the underside is 

M b 762000 onrn .. 
a b =—= = 2060 lbs. per sq.m. 

In the examples so far considered it was assumed that the base plate was a frame 
open in the center, i.e., consisting of the lateral beams connected across the ends by- 
two other beams carrying the bearings. This is the usual construction for small and 
medium-sized machines. In large-size vertical machines this base-plate frame is usually 
closed on the under side by giving it a curved bottom to fit the crank circle. In 
that way the four sides of the base plate are connected in such a way as to materially 
increase the strength and stiffness of the entire plate. This is shown more clearly by 
the following example, which shows a computation not only for the section of the 
plate directly under the bearing, but for the entire cross-section. The fact that the 
bed plate is not symmetrical with reference to the cylinder axis, compels us to con¬ 
sider the entire section. Had it been symmetrical, only one half of the section would 
need to be considered. 

I 

Example. 4. Cylinder diameter D = 19.7", maximum total pressure on piston=280 XI9.7 2 = 
109000 lbs. Distance between points of application of force (see Fig. 50), a = 49.3". Bending 



moment due to the total piston load (not half this load, since the entire cross-section is 
considered) is 


M b = 


109000 X 49.3 
4 


= 1338000 in.-lbs. 


The moment of inertia of the cross-section is to be determined exactly according to the dimen¬ 
sions of the actual construction shown in Fig. 66. This makes it necessary to subdivide the 
section into a number of partial areas. Many of these partial areas, as I, IV, and VI, could 
without great error be entirely neglected, others with minor changes could be much simplified. 
Nothing of this kind has been done in order to show the complete method of computation for 
a definitely given cross-sectional area. 

Cross-section through left hearing: 

59X1 77 s 

h = 1 —^— + (.59 X 1.77 X4.65*) = .27 + 22.6 = 22.87 in. 4 


1 11 = 


1.3X1.02* 


+ (1.3X1.02 X 4.25 2 ) = .12 + 24.0 = 24.12 in. 4 


12 














































GENERAL ENGINE PARTS 


99 


, .78 XI.58 3 

hu =-—-+ .78 X1.58 X 4.57 2 ) = .26 + 25.7 


/iv = 2 


12 
.39 + .78 


2 X2 - 36 ’ /^ 9 +£ 8 x23e><641 


12 


• + 


=25.96 in. 4 


= 2(.64+57.0) =115.28 in. 4 


1 97 X 59 3 

/v = —■ • + (1.97 X .59 X 4.05 2 ) = .034 + 19.10 


/vx = 2 


12 


.71X 


.98 +1.46\ 3 


+ 


12 

87X17.6 3 


•+( .71X——t—-X6.5 : 


= 19.13 in. 4 


= 2(.ll + 36.6) =73.42 in. 4 


•Ivin — 


12 

10.9X2.06 3 

12 


•65 2 ) j =: 


+ (.87 X 17.6 X1.65 2 ) =2(395 + 41.8) 


+ (10.9 X 2.06 X 6.1 2 ) = 7.7+829.0 


to QY 71 3 

Iix = _• -- - + (10.9 X .71X 1.26 2 ) = .32 +12.3 


Ix = 


/xi — 


12 

10.9 X.78 3 
12 

2.13 X.87 3 
12 


+ (10.9 X .78 X 10.1 2 ) = .51 +862 

+ (2.13 X .87 X 3.96 2 ) = .12 + 29.3 


63 V1 38 3 

/x n = ^ - + (.63 X 1.38 X 4.65 2 ) = .13 +18.8 


For the crank case 

23.6 X.86 3 


=873.60 in . 4 


=836.70 in . 4 


= 12.62 in. 4 


=862.51 in. 4 


=29.42 in. 4 


= 18.93 in. 4 


3/^ = 2914.56 in. 4 


/xm — ■ 


12 


+23.6 X .86 X10 2 = 1.34 + 2030 = 2031.34. 


' Cross-section through right main bearing. 

On this side the areas XIV, XV, and XVI appear instead of the area VIII of the left 
section. All other areas are exactly the same on both sides. Hence for the right section we 
have simply 


->[- 




44X2.06* 
12 

03X3.46* 


+ (2.44X2.06X6 


/xvi — 


12 

3.95 X.98* 
12 


+ (1.03X3.46X5 


•l 2 ) j =2(1. 

.4*)] =2(3. 


78 + 188.0) =379.56 

54X104.0) =215.08 


+ (3.95 X .98 X 3.86 2 ) = .31 +58.0 


58.31 


Total =652.95 


























100 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

Hence moment of inertia of right cross-section 

I r = 2914.56 -836.70 + 652.95 = 2730.81 in. 4 

which is 183.85 in. 4 smaller than for the left side. 

From the partial moments of inertia the total moment for the entire section is 

/ = h + / XI1I + I r = 2914.56 + 2031.34 + 2730.81 = 7676.71 in. 4 

Since the distance from the neutral axis to the fiber under greatest stress is «- 10.45", the 
section modulus will be 


Tf 


t _K76J1 
10.45 


and finally, from the value for Mb above computed, the maximum fiber stress is 

1 338 000 


Ob — 


735 


= 1820 lbs. per sq.in. 


The moment of inertia of the crank case amounts to about 35% of the sum of the 
moments of the two bearing cross-sections. If these last two are considered alone, the 
total section modulus will be only about 484, since the fiber distance is increased to 




Fig. 67.—Dimensions in Inches. 


Fig. 68 .—Dimensions in Inches. 


about c = 11.8", leaving out the crank-case wall. It is evident, therefore, that this 
wall, or the lower surface of the base plate, increases the strength of the casting very 
materially. 

As a general thing, the fiber stress in the middle vertical section of the base plate 
perpendicular to the axis of the crank shaft is less than in the vertical sections 
through the axis of the bearings. Under certain circumstances, however, as for 
instance great distance between bearings, bad connections between fram'e columns and 
the plate, the reverse condition may exist, so that a check computation should be 
made. The following example shows the method. 

Example. 5. Suppose a base plate similar to the one preceding is to stand a piston load of 
Pz = 220000 lbs., corresponding to 110000 lbs. pressure on each bearing. Let the distance 
between middle sections of bearings be 35.5". The bolts connecting columns and plate are 

220000 

distributed in five groups, according to Fig. 67, so that each group is loaded with - 
= 44000 lbs. 

The first thing to do is to determine the resulting bending moment due to P z distributed 
in the planes marked I to V. The best means of doing this is the force polygon. Since the 
component forces are of equal magnitude and are distributed symmetrically about the middle 
plane III, the resultant force will be in plane III, and it is necessary to consider or make the 
computation for only one half of the plate, as is done in Fig. 68. From this the maximum 






















GENERAL ENGINE PARTS 


101 


bending moment Mb re s. = 1005000 in.-lbs. (In explanation, Mb Tei =HbKl in in.-lbs., when H = polar 
distance in the force polygon, b is the maximum distance cut off by the two closing lines of 
the force polygon on the line of the resultant force, K = scale on which the forces are repre¬ 
sented, and i = scale of length. Fig. 68 
shows the force polygon one fourth of 
t^e original size. In this original draw¬ 
ing tf = 5.90", 5 = 2.44", K = 27 940 lbs. 
per in., and 1= 2.5" to the inch. Hence 

M b re3 . = 5.90X2.44X27940X2.5 
= 1005000 in.-lbs., 



as given above.) 

For the determination of the section 
modulus of that section of the base plate 
under greatest stress, perpendicular to 
the crank shaft, the dimensions are given 
in Fig. 69. As indicated there, the total 

area must again be divided into a number of partial areas. The left half of the section, not 
shown, is the same as the right. As before 


h — 2 


.59X.75X.37 2 ) 


= 2(.0210 + .062) =1.66 in.‘ 


/ lI = 2^^^--- + (1.02X8.25X4.13 2 )j =2(47.7 + 144) =383.40 in. 4 
7m = 2——* 2 -- 53 + (12.2 X 2.95 X 9.85 2 ) J =2(26.1 + 3490) = 7032.20 in. 4 

/iv = 2 + < 19 - 7 X 1.02X.51 2 ) j =2(1.75 + 5.27) = 14.04 in. 4 

For area V, considered as half of a circular ring, the general formula for the moment of 
inertia about its own neutral axis is 


/ = .1098(£ 4 —r 4 ) - 


.283 Rb'iR-r) 
R+r 


and, with the dimensions of Fig. 69, we shall have 

.238 X 23.6 2 X 22.8 2 (23.6 - 22.8) 
I = .1098(23.6 4 - 22.8 4 )- 


= 3090 in. 4 


23.6 + 22.8 

The distance from the center of the semicircular ring to its neutral axis is 


_4_ /fl* 


-r’\ 

_ 4 , 

/23.6 s -22.8 3 \ 

■rV 

3?r 

I23.6 2 —22.8 2 / 


= 16.85". 


The position of this neutral axis is therefore 


16.85 - (23.6 -15.25) =8.5" 























102 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

below the neutral axis o-o of the section in Fig. 69. Hence the moment of inertia of area V 
with reference to axis o-o is increased to 


ly = 3090+8.5 2 X 2 x (23.6 2 - 22.8 2 ) = 3090 + 3540 = 6630 in 4 . 

The sum of the partial moments gives the total moment of inertia for the section 
I = 1.66 + 383.40 + 7032.20 +14.04+6630 = 14061.3 in 4 . 

Since e x the maximum fiber distance, = 15.25", the section modulus is 


TF = 


14061.3 

15.25 


= 922.0. 


The bending moment, Mb T ea. = 1005 000 in.-lbs., above determined, therefore, gives a maxi¬ 
mum fiber stress in the section examined of 


1 005 000 
922.0 


= 1090 lbs. per sq.in. 


It is clear from the above that in the central plane considered the stresses may under certain 
circumstances be quite considerable. 


3. Engine Beds. The old type engine beds with overhung cylinders are being 
replaced more and more by a type of frame in which the cylinder is entirely, or at 
least partly, supported by the frame. For equal weight this furnishes a construction 
of much greater stiffness and strength with pleasing form. For that reason these 
frames are now also used extensively for small machines. For large machines the two 
side walls of the frame are usually connected by a crank-case wall across the bottom, 
the whole being cast in one piece. In the very largest models, however, the side 
walls are often separately cast. 

Single-beam frames, so-called bayonet frames, which as Corliss or Tangye frames 
have found application up to the highest powers in steam-engine practice, have not 
shown themselves well adapted to gas-engine construction. Delamare and Deboutte- 
ville tried this type of engine bed in some of the first forms of their large simplex 
engines, but were compelled to abandon it very soon. The reasons for their failure 
are not far to seek. In a Corliss frame, the main crank-shaft bearing must take care 
of the entire piston load, and since this load acts at an arm equal to the distance 
between center line of cylinder and center line of bearing, the bearing pressures are 
nearly twice as great as in the case of other frames. On the other hand, the double 
beam, or center crank type of frame, distributes this load between two bearings, and 
since, in this case, the moment arm above mentioned does not exist, the load on each 
bearing is simply one-half of the total piston load. Further, the Corliss frame, owing 
to the manner in which the load on the piston is taken up, is subject to a consider¬ 
able bending moment in the horizontal plane. This is not found in the ordinary type 
of frame. In Corliss frames, therefore, not only the crank-shaft itself, but also the 
frame parts must be made considerably heavier. These and other disadvantages, make 
this frame unsuited for, or at least little adapted to, the larger sizes of gas engines. 
This is apparently contrary to the experience of steam-engine designers, which, how¬ 
ever, owing to the different conditions, has no bearing on this case. 1 


1 Translator’s Note. To judge from present American practice with regard to large engines, the 

difficulties mentioned have been largely overcome. 







GENERAL ENGINE PARTS 


103 


Constructive Details. The height of center of cylinder and bearings above the floor 
should, on account of the work of starting, lubricating, etc., not be less than about 
28" in the smaller engines; in large machines, from consideration of strength mainly, 
it should not exceed from 48-60". For the beds of the smaller machines, the general 
shape of the casting and the requirements of foundry practice usually call for a mini¬ 
mum thickness of the walls of the hollow casting such that stresses exceeding the 
allowable hardly ever occur. Of so much the greater importance, however, is the 
careful design of large frames if sufficient strength and stiffness are to be obtained in 
suitable form and without waste of material. 

The dangerous section in horizontal engine beds, the section which must be most 
carefully considered in design, is usually found in a plane perpendicular to the cylinder 
axis, and located somewhere between the main bearings and the seat of the cylinder, 
see Fig. 70. In this plane, the bending moment, whose arm is the distance from the 



Fig. 70.—Result of Poor Frame Construction. 

center line of the cylinder to the neutral axis of the frame section under considera¬ 
tion, and which moment adds itself to the tensile stress already acting on this section, 
reaches it's maximum. Other parts of the frame under severe stress are the vertical 
sections through the middle planes of the main bearings, especially on the fly-wheel side, 
where, besides the piston load, gravity forces also come into action. These, together 
with the cylinder flange, all require careful checking. An increase in the capacity of 
an engine not only calls for an increased cylinder diameter, but with it also increase 
the height of the frame and the magnitude of the bending moment above mentioned. 
For this reason the difficulties encountered in economic but safe frame design, increase 
very rapidly as the upper limits of capacity are approached, especially since at the 
same time perfection of workmanship in the shop grows less certain. The constructive 
means which the designer may employ in such cases will be shown in later examples. 

The necessity of keeping the most important bending moment with which we have 
to deal in the design of a frame down to its lowest possible figure, calls for a design 




104 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


of the vertical cross-section concerned, such that the neutral axis of this section shall 
be as close as possible to the center line of the cylinder. It is radically wrong, there¬ 
fore, to mass great bodies of metal near the bottom of the section; the metal is 
needed in the upper parts of the section. Fig. 70 shows a good example of what is 
likely to happen when this simple rule is neglected. Further, the height of the crank 
bearings should be. kept as low as possible, while it is good design to carry the lateral 
walls as high as they can be made. For small engines, these walls should extend at 
least to the horizontal center plane of the cylinder; for large engines they should go 
beyond this plane—the higher, the better. It is a costly mistake, and not one pleas¬ 
ing to the eye, to keep the side walls of the frame much below the level of the crank 
bearings, although even to-day this is quite often done. Gertain forms of main bear¬ 
ings (as for instance, those shown in Figs. 71-80, inclined inward) compel this statically 
wrong construction of the side walls of frames, and their use is, therefore, permissible 
only in the smaller sizes. 

The most central taking up of the force of the explosion in the cylinder is a 
necessity in gas-engine construction, and this is a fundamental rule which can not be 
too often repeated. Its observance even in the design of small engines is of advantage. 
Unfortunately, the rule can only be partially followed in the usual types of frames. 
Even in the most favorable cases, the distance between the neutral axis of the 
dangerous section and the center line of the cylinder is sufficiently great to result in 
a bending stress approximately equal to the tensile stress on the section, so that the 
total stress is nearly double what it would be if the load P z could be balanced 
exactly in a central line. If the common type of frame is abandoned, an approxi¬ 
mately central distribution of forces can be obtained most easily by putting in, above 
the horizontal central plane, a rigid connection between the main bearings and the 
seat of the cylinder. See Figs. 83-86. This idea was already carried cut in the older 
Simplex engines; the latest models still retain it, and rightly so. 

Designs of Horizontal Frames. 



Figs. 71-74.—Frame for a 6 H.P. Gas Engine. 







































































































GENERAL ENGINE PARTS 


105 




Figs. 81 and 82.—Frame with Cross-head Guide for a Diesel-Giildner Engine. 

The lower part of the bored guide is water cooled throughout the contact length. For details of the 
main bearing, see Figs. 109 and 110, p. 115. (The gear case a communicates with the oil chamber of the 
main bearing in order to increase both the quantity of the oil in circulation and the external cooling 
surfaces. After a slight turning of the cross-head, the wrist-pin may be loosened through the opening b.) 



Fro. 83.—Frame for a 700-800 H.P. Simplex Engine, Old Type. 





































































































































































































































































| < 0 * 405 ^ 


106 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 





























































































































Figs. 87-89 — Frame for a 300 H.P. Gas Engine. L. Soest & Co., Diisseldorff. 



Figs. 90 and 91.—Frame for one of the Older Type of Nurnberg Gas Engines, 160 H.P. 

(After the piston rod is removed, the piston may be pulled forward to the position b, resting on the ledges a. The 
niston is then easily taken out without disturbing the connecting rod. c is a stuffing-box in the pipe system for cooling 
the piston The cylinder-oil drip gathers at d and e. The level in d is maintained to the same point, the oil serving to 


lubricate the slipper cross-head.) 


107 


































































































































































































































































































































108 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Examples. 1. The dangerous section in the frame shown in Fig. 92 is to resist a maximum 

load of 17600 lbs. The area of the section = 37.3 sq.in., 
hence the tensile stress in the section is 



Fig. 92.— Dimensions in Inches. 


Pz 

a= T 

As before, the 
tion is found from 
follows: 


17600 

37.3 


= 470 lbs. per sq.in. 


total moment of inertia of the see¬ 
the moments of the partial areas as 


h = 1,18 X — + (1.18 X .59 X 7.58 2 ) 
1 2 


40.02 in. 4 


/ n = -' 8 * 103 +(.78X10X2.3 2 ) =106.30 in . 4 

I IU = 6,10 X - 6 + (6.1 X .67 X ,92 2 ) = 3.59 in . 4 

X 2 

47 V 71 3 

7 IV =—-+ (.47X.71X4.1 2 ) = 5.61 in . 4 

-L w 

/v= 3^4>C8D + (3 94 X 87X4 Q 2 ) = 82.21 in . 4 

12 

Q7V I QQ3 

I V1 = - — + (.87X1.33X3.33 2 ) = 12.97 in . 4 

12 

1 C7V AQ3 

Ivn = \w~— + (1 -57 X .43 X 4.25 2 ) = 12.21 in . 4 

For ^ of the cross-section, 2 (/ n ) =262.91 in . 4 
and since e! = 7.87", the section modulus is 


2X262.91 
4 = 7^87 


66 . 8 . 


Distance between neutral axis of section and center line of cylinder (CM) =c = 6.5". Hence 
maximum bending moment 

476 = 6.5X17600 = 114300 in.-lbs. 
and therefore bending stress in the outer fiber is 

114300 i-mm 
<76 = - ^ ■ =1/10 lbs. per sq.m. 

The maximum total stress in the fiber is consequently 


<t + <76 = 470+1710 = 2180 lbs. per sq.in. 



























GENERAL ENGINE PARTS 


109 


In spite of the fact that in this case the frame walls extend If" above the center 
line of the cylinder, the bending stress is more than 3f times the tensile stress. 

2. The frame section under greatest stress, shown in Fie. 93, 
is for an engine for which Z) = 13.75"; the total area of the section 
is 107 sq.in. At the inner dead center position of the piston the 
load will be P z = 280 Z ) 2 = 280 X13.75 2 = about 53000 lbs. The tensile 
stress in the section will be 


53000 inr 1U 

a =-— = 495 lbs. per sq.m. 

107 


The bending moment, due to Pz and the arm c = 10.2", is 
Mb = 53000X10.2 = 540 000 in.-lbs. 

Moments of inertia of the partial areas, referred to the neutral axis: 


r. 7 v S7 3 

1 1 = + (5.7 X .87 X11.9 2 ) = 703.3 in . 4 



Fig. 93. —Dimensions in 
Inches. 


In — 


Illl — 


/lV = 


Iv = 


I\ I- 


12 

.71 X 15 3 

12 

.79X17.3 3 

12 

.75X2.76 3 

12 ~ 

3.34X1.18 3 
12 

15X.79 3 


+ (.71X15 X 3.94 2 ) = 365.0 in . 4 

+ (.79X 17.3X2.76 2 ) = 443.0 in . 4 
+ (.75 X 2.76 X 4.92 2 ) = 51.3 in . 4 

+ (3.34X1.18X6.9 2 ) = 188.5 in . 4 


/v II — 


12 

8.7X.79 3 

12 


+ (15X.79X6.7 2 ) = 530.6 in . 4 

+ (8.7X.79X1.3 2 ) = 11.9 in . 4 


Moment of inertia for 
entire section 
= 2X2293.6 = 4587.2 in 4 . 


Sum of moments for half section = 2293.6 in . 4 
Since e! = 12.0", the section modulus is 

^ = 458^2 = 3830 

1 z 


and the greatest bending stress in the upper fiber will be 


540 000 
ab ~ 383 


= 1410 lbs. per sq.in., 


and hence total stress in the outer fiber of the section is 

^ + ^=495 +1410 = 1905 lbs. per sq.in. 





































110 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


3. For a Niirnberg engine, cylinder diameter = 35.4", the section under greatest stress in 
each one of the beams forming the side walls of the frame is shown in Fig. 94. The explosion 
pressure is as high as 450 lbs., so that the maximum load on each beam will be 



Fig. 94.—Dimensions 
in Inches. 


p _.78ax3M»x45o_ lb3 . 


The area of section is 260 sq.in., so that the purely tensile stress is 

221000 


a = 


260 


-=850 lbs. per sq.in. 


Distance between neutral axis and cylinder axis is c = 7.9", so that the 
maximum bending moment is 

Mb = 7.9 X 221000 = 1740 000 in.-lbs. 


Moments of inertia of the partial areas (the distances of the center of gravity of each area 
from the neutral axis of the section are not given in the figure): 


18 5X3 15 3 

h = ■ -—-+ (18.5X3.15X 14.9 2 ) = 12 998 in. 4 

Ji^i 

9 06 X13 4 3 

/n-—— • + (2.06X 13.4X7.3 2 ) = 1 883 in. 4 

_ 2.16X13.4 3 

Ini—' -—-•+ (2.16X 13.4X7.3 2 ) = 1 968 in. 4 

r 2.16X13.4 3 / 

/iv =-—-+ (2.16X 13.4X7.1 2 ) = 1893 in. 4 

1 38 v 11 ft 3 

/v = — 12 + (1.38X11.8X5.9*) = 754 in. 4 

1 38 X 21 8 3 

/vi=-— ‘ + (1.38X21.8X 10.8 2 ) = 4690 in. 4 

j 1.57X21.8 3 , 

Iy n=- ^2 -+ (1.57X21.8X10.8 2 ) = 5 390 in. 4 

1.38X5.9 3 

Ivm = -——- + (1.38X5.9X 14.8 2 ) = 1808 in. 4 

r 1.38X4.13 3 , 

/ix= -12-+ (1.38X4.13X19.8 2 ) = 2 248 in. 4 

. 11X1.77 3 

/x== - 12 - +(11X1.77X22.7 2 ) =10005 in. 4 

= 43 637 in 4 . 


Moment of inertia of 
section of side wall 
2 (/ n ) =45614 in . 4 


Distance of upper fiber from neutral axis of section is 13.4 + 3.16 = 16.56". Hence section 
modulus is 

































GENERAL ENGINE PARTS 


111 


and maximum bending stress will be 


(76 = 


1740000 

2630 


=660 lbs. per sq.in. 


Total fiber stress is therefore 

(7 + 06 = 850 + 660 = 1510 lbs. per sq.in. 


From this example the good effect of proper design of cross-section upon the fiber stress is 
clearly seen. In spite of the large cylinder diameter, the distance from the center of the cylinder 
to the neutral axis of the section is only 7.9", and the bending stress is in this case not larger 
than the tensile stress, as is commonly the case, but is about one fourth 


smaller. 

4. In the Niirnberg blast-furnace gas engine, shown in Plate XVIII, 
cylinder diameter 52", the dangerous section of the frame is that shown 
in Fig. 95. Each beam constituting the side walls is cast in two 
pieces, and the soft steel tie-rod XIII, Fig. 95, placed in the neutral 
axis of the section serves, in the first place, to unite the parts of the 
side wall, and in the second place relieves the castings of some of the 
stress. 

Assuming the explosion pressure to be 430 lbs., the maximum 
load on each beam of the frame will be 


P = 


430X.785X52 2 

9 


= 455 000 lbs. 





Fig. 95.—Dimensions in 

. . . Inches. 

It is not known how this load is distributed between the casting 

and the tie-rod; it is therefore necessary to make some kind of an 

estimate. Assume that the tie-rod is subjected by the load P to a stress equal to 4250 lbs. per 
sq.in. In that case the purely tensile stress on the cast-iron section, whose area is about 776 
sq.in., will be 


455 000 - (.785 X 7.9 2 X 4250) 
776-48.3 


= 346 lbs. per sq.in. 


Moments of inertia of the partial areas referred to the neutral axis of the entire section: 

23.6X3.96 3 


7i = 


/ n = 


I in — 


7iv = 


/v = 


/v — 


12 

2.46X9.45 3 

12 

2.96X9.45 3 

12 

2.96X9.45 3 


12 

1.97X51.2 3 
12 

2.56 X 57.5 s 


+ (23.6X3.96X47.3 2 ) =209120 in.* 
+ (2.46 X 9.45 X 40.6 2 ) = 38 370 in. 4 
+ (2.96 X 9.45 X 40.6 2 ) = 46 300 in. 4 
+ (2.96X9.45X40.6 2 ) = 46300 in. 4 
+ (1.97 X 51.2X 10.05 2 ) = 32 200 in. 4 
+ (2.56X57.5X7.1 2 ) = 48030 in. 4 


12 


























112 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


1 Q7xll 6 3 

/vu= - + (1.97X11.6X12.60 = 3880 m. 4 


/vm — 


12 

1.97X15.95 3 

12 

1.97X27.6 3 

12 

2.56 X 21.3 3 
12 


+ (1.97X 15.95X14.8 2 )= 7540 in. 4 

+ (1.97 X27.6X30.3 3 ) = 53440 in. 4 
+ (2.56 X 21.3 X 32.3 2 ) = 58 760 in. 4 


19 6X2 36 s 

Ixi = ' V ' -+ (19.6X2.36X44.1 2 ) = 90020 in. 4 

/xn = .05(13.8 4 —8.86 4 ) 
ttX7.9 4 


7x III- 


64 


= 1500 in. 4 

190 in. 4 


Moment of inertia of entire cross-section = 635650 in. 4 

including the tie-rod. 

After subtracting partial moment / X m there remains for the cast-iron cross-section 

7 = 635 650-190 = 635 460 in. 4 
Hence the section modulus, with ^ = 49.2" 

I 635460 

= W = -= _ =12920. 


49.2 


The section modulus of the tie-rod is 


190 

3.94 


= 48.2. 


The maximum bending moment on the entire section, with c = 33.8", is 

Mb = 455 000 X 33.8 = 15400 000 in.-lbs. 

Since it is assumed that the tie-rod is put under a stress of 4250 lbs. per sq.in., by the load P 
the part of the above moment Mb falling to the rod will be 

4250X48.2 = 205 000 in.-lbs. 

The maximum bending moment for the cast section alone is therefore 

Mb = 15400000-205 000 = 15 195 000 in.-lbs. 
and hence the bending stress in the outer fiber is 


15 195 000 „ 

° b -lb yop— = 1175 lbs. per sq.in. 











GENERAL ENGINE PARTS 


113 


The maximum total stress on the outer fiber of the cast iron is therefore 

o+ob'= 346 + 1179 = 1525 lbs. per sq.in. 

The maximum stress in the outer fiber of the steel tie-rod will be 

4250+4250=8500 lbs. per sq.in. 

To this must be added the constant but unknown stress induced in the. rod by the drawing up 
of the nuts. This stress, however, is of comparatively small importance. 


In case the cylinder is overhung, the connecting flange in the frame must be made 
abundantly wide and strong, because this flange, besides being loaded by the explosion 
pressure P z is also subject, during the expansion stroke, to the lateral thrust N of 
the connecting rod, which may under certain circumstances cause a considerable 

p 

bending moment in the flange. The maximum value of N is approximately y^, and 

this maximum occurs at from 15-18%- of the stroke. Consequently, the point of 
application of AT max is much nearer to the head of the cylinder than to the connecting 
flange, especially if the piston pin is well toward the face of the piston, and the 
moment arm on the flange is correspondingly long. For this reason the overhung 
cylinder should be used in small engines only. If, however, the cylinder proper is 
merely a liner inserted into a barrel extension of the frame casting, the barrel forming 
the jacket wall of the cylinder, this jacket wall must take care of the load due to Pz, 
which is always considerably in excess of the bending stress due to A max . The latter 
is taken care of by both the jacket wall and the liner. For further details on these 
points, see “Cylinders,” p. 129. 

6. The best form for the main bearings for horizontal engines is the one in which 
the explosion load is taken up directly by the frame. This is possible only when 
the dividing line of the bearing slants inward, Figs. 71-80, or is horizontal. Bearings 
slanting outward, Fig. 107, are not suitable because the maximum load comes on the 
caps and bolts. Further, the journals run upon the lower dividing line of the bearing 
all the time, because the influence of the weight of the fly wheel upon the direction of 
pressure on the journal is of small importance. In view of these disadvantages, the 
good points that this type of bearing may have are decidedly overbalanced. The most 
suitable form of main bearing is undoubtedly that slanting inward, for the explosion 
load is always taken up by the frame body directly and only the lower half of the 
bearing is ever under load. But this form requires low frame walls, and since these 
are not suitable for the larger models, the proper use of this type of mam bearing is 
necessarily confined to the smaller machines. For the large horizontal gas engine there 
remains only the horizontally divided main bearing. 

In vertical machines, with crank shaft below the cylinder, the journal pressure due 
to Pz always falls to the lower half of the bearing. It should be noted, however, 
that the weight of the fly-wheel in this type of engine causes an upward pressure 
on the cap of the opposite bearing when the piston is unloaded. This pressure is, 
however, never great enough to determine the dimensions of either the cap or the 
bolts. The last statement also holds for vertical frames with shaft above the cylinder. 
In this type of frame the direction of the journal or piston pressure is always through 
the cap of the bearing, whose dimensions are to be determined accordingly. 


114 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


The lubrication of main bearings 
should be thorough, certain, and 
easy of inspection. Drop-feed oilers 
are suitable for small machines only. 
For medium-sized engines ring oiling 
is much in favor, although this 
method also has many weak points 
(Certain cutting down and weaken¬ 
ing of the bearing surfaces, opera¬ 
tion not directly open to inspection, 
gradual deterioration of the lubricant, 
etc.) For the bearings of large 
engines, forced lubrication by means 
of special oil pump is the best 
method. 


Designs of Main Bearings. 




Figs. 98-101.—Main Bearing, 12 H.P. Loutzki Engine. 



Figs. 102 and 103.—Main Bearing for a 20 H.P. Bdnki Engine, Cyl. Dia. 10", Stroke 16". 








































































































































































GENERAL ENGINE PARTS 


115 


Figs. 104-106.—Main Bearing for an 
80 H.P. Vertical Diesel Engine, 
dia. 16", stroke 24". Built by 
Maschinenfabrik, Augsburg. 

The gears driving the intermediate 
shaft are located in the center-plane 
of the left hand main bearing. 

(This construction was originally 
used by Loutzki.) 



Figs. 107 and 108.—Type of Bearing 
used on Korting Engines of from 
2-16 H.P. 



Figs. 109 and 110.—Bearing for Frame shown 
in Figs. 81 and 82. 


Constructive Details. For diameter and length of bearing, or allowable bearing 
pressures, see Crank-shafts, p. 160. 

The material for bearings up to 5" journal diameter (d) is nearly always bronze, in 
which case the thickness of the metal may be made 

s = (j^ + -20^ inches.(4) 

For larger journal diameters the bearings are made of cast iron, lined with white 
metal, Figs. Ill and 112. The least thickness of the shell of the bearing proper 
(without the lining) may be made about 

s'=^^ + .4^ inches, 


(5) 
























































































116 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

In either case the collars at the end of the bearing may be given, see Figs. Ill and 

112 , 

a radial height, ' .8 to 1.1 s or s', ] ^ 

and an axial thickness, 6 = 1.0 to 1.2 s or s'. J 

If special circumstances make it necessary to keep the external diameter of the 
bearing as small as possible, cast steel or malleable iron may be used in place of cast 
iron. In any case a white metal liner is employed. The thickness of the walls s 
may then be taken according to eq. (4), the same as for bronze. 

On account of sudden variations of pressure, the proper choice of the white metal 
alloy 1 and the manner of holding it in place are of the greatest importance. The 
shell of the bearing should be furnished with sufficient dove-tail grooves both circum¬ 
ferentially and longitudinally, Figs. Ill and 112, and the surfaces should be well tinned 

over before the lining is cast in. If these precautions 
are not taken, the lining is. apt to be squeezed out of 
shape and loosened in the shell. Care should also be 
taken that the ends of the bearings, where they may 
come in contact with the crank shaft, taking up any 
lateral pressure that may occur, are well covered with 
white metal. 

For reasons of economy in manufacture, the outside 
surface of the bearings is without any projection, so 
that it can be finished in the lathe. This permits the 
bearing to rest evenly over its entire length in its supports. The shifting or turning 
of the bearing in its seat is prevented by safety pins or plates afterward inserted or 
screwed on (see x, Fig. 112). The lower half of the bearing is not thus secured in 
order to be able to take it out, in case of necessity, by slightly raising the shaft and 
turning the half bearing about the shaft until it comes on top. 

In the larger horizontal machines lateral adjustment in the bearings may have to 
be provided for. In such a case the adjusting mechanism should always be placed 
in the side of the bearing nearest the cylinder in order to protect the adjusting screws 
or wedges against the full load on the bearing. Three part bearings, Figs. 111-112, 
may in such cases be successfully employed. The adjusting wedge, y, and the movable 
block, z, are placed in the side of the bearing under least stress, so that all joints are 
out of the direct line of journal pressure. 



Examples. 1. The main bearing, Fig. 107, slanting outwards at an angle a =40° with the 
vertical, is to carry a load of P = iP* = 9900 lbs. The component P n of the force P, passing 
normally through the cap of the bearing will be 

Pn=P' sin a =9900 sin 40° = 9900 X .643 = 6380 lbs. 

T his is the load which determines the dimensions of the caps and bolts by the method of 
Example 1, p. 89. The load on the bearing furthest from the fly-wheel is relieved somewhat 
by the weight of the wheel, but an additional load is often caused by the pull of the driving 


1 Good alloy for heavily loaded main bearings: 8 parts by weight of copper, 12 of antimony, and 
80 of tin. 

















GENERAL ENGINE PARTS 


117 


belt, and hence this should not be taken into account. Pre-ignition also causes an increase in 
P n , and for that reason, too, a good factor of safety is essential. 

2. The main bearing shown in Fig. 113 slants inward. The load P at the moment 
of explosion is 11000 lbs. This horizontal force causes a shearing 
stress in the section x-x whose area is 11 sq.in., of 


^ = 11000_ lOoo lbs. p er sq.in. 



Fracture, however, would in all probability not take place in the sec¬ 
tion x-x, but in y-y. This dangerous section must therefore also be 
examined. Any influence that the cap may have to relieve the load 
on this section should always be neglected. The component Pn, act¬ 
ing at the neutral axis of the section y-y is equal to Pn = P sin a'= 11000 X sin 20° = 3760 lbs. 

The area of the section, Fig. 114, is 12.6 sq.in., hence the tensile 
stress is 



3760 90" IK 

a= =29/ lbs. per sq.m. 

12.6 


Besides this there is acting on this section a bending moment 
Mb = PI = 11 000 X1.65" = 18100 in.-lbs. 


'*—"■334 -> 

Fig. 114.—Dimensions in 
Inches. 


The cross-section y-y is shown in Fig. 114. 
moment of inertia 


From this the 


7 = /i + /ii + /iii + 7iv — 


4.72X1.57 3 

12 


+ (4.72X1.57X1.1 3 ) 


+ 2 


.47X2.44 3 

12 


+ (.47 X2.44X .71 2 ) 


+ 


3.94 X. 71 : 
12 


+ (3.94 X.71X 2.48 2 ) 


= 30.2 in. 4 


The maximum bending stress is therefore 


Ob 


Mbe 2 IS 100X1.89 u 

. + = _x-= 1130 lbs. per sq.m. 


$ - 


30.2 


The combined stress. 0 + 06=297 + 1130 = 1427 lbs. per sq.in., would be much smaller if the 

The d followingTs al a S s e y!tematic investigation made by Professor Bach on a fractured main 
bearing of a steam engined It serves to show very clearly to what extent the safety of a con¬ 
struction sometimes depends upon the judgment of the designer, and how, in the case of mam 
bearings, the usual methods of computation are sometimes badly misleading. 


1 Zeitschrift d. V. D. I., 1901, p. 1561. 































118 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Fig. 115 shows the fractured pedestal, one break, a-b, starting from the corner of the seat 
of the bearing and extending diagonally toward the base, the other, c-d, is in the cap. To 

obtain some idea of the load at rupture by actual measurement, cast-iron test specimens, 
Fig. 116, were made in which the dangerous section, under stress due to the force P, had 

dimensions as close as possible to those of the actual sur¬ 

face of the break in the pedestal. These specimens broke, 
through the section indicated in Fig. 116, under a load 

P = 46700 lbs. The dimensions of the cross-section are given 
in Fig. 117. 





Fig. 116.—Dimensions in 
Inches. 


.65 


Fig. 117. —Dimensions in 
Inches. 


The common method of computation, using the dimensions and data given in the figures, 
then gives the following results: 


P = 46700 lbs., 


a = 40°, P„ = 46700 sin 40° = 29 000 lbs., 

(7.68 X 4.58 X 2.29) - (5.72 X 2.75 X 2.56) 
19.45 


a = 2.92 in., 
= 2.07 in., 


c = 1.23 in., 


in which 19.45=/= area of the cross-section. Moment of inertia of the section of the break 


is 


/ = 


7.68X4.58 3 _ 


12 


+ 7.68X4.58(2.29-2.07) 5 


5.72X2.75 3 

12 


+ 5.72X2.75(2.56 —2.07) : 


= 50 in. 4 


The section modulus is therefore 


wJ 7-W~= 2i - 2 - 


from P 


Finally, the stresses involved are 

Pn 29000 


n, 




f 19.45 


= 1490 lbs. per sq.in., 


<>v\ 


and from P, 


Ob = 


P(a + c ) 46709X4.15 


IF-; 


-= 8000 lbs. per sq.in. 


• (D 


Test specimens of the same cast iron under transverse tests showed a ratio of strength in 
the outer fiber of Kb:K z l = 1.79:1, and a strength in the outer fiber of X& = 33 300 lbs. per sq.in. 


1 See Bach, Elasticitat and Festigkeit, 2 eel., eq. (77) or Maschinen Elemente, 8 ed., eq. (137) 









































GENERAL ENGINE PARTS 


119 


The above computation, based on actual tests, however, shows a total stress in the fractured 
section of only 


1.79a + ob = (1.79X1490) +8000 = 10700 lbs. per sq.in., 


10700 

which is 1 “ 3330 Q - 


two-thirds smaller than the actual 


strength of the material in the outer fiber. 


The above method of computation therefore seriously underestimates the actual existing stress, 
and the reason is that the outside fiber, which is under greatest stress has been considered as 
straight, while in reality, owing to the fillet put in with a radius equal to .31", it must be 
considered as an element with considerable curvature. The moment due to P tends to straighten 
out this fiber, and of course intensifies the stress very materially, as may be shown as follows. 

Since the radius of the curvature of the fillet is .31", the mean radius of curvature for the 
neutral plane of the section will be r = .31+e = .31+ 2.07 = 2.38". The stress induced in curved 
rods, according to Bach, 1 is 


Pn Mb M b e 
f fr xfr r + e 


lbs. per sq.in., 


( 8 ) 


in which, from the above figures, 1+1 = 29000 lbs., /= 19.45 sq.in., Mb — —46700X (2.92X1.23) 

= -193500 in.-lbs., 2 and x = \ f— —d/=.844. 3 

/ J r + e 

The maximum stress in the outer fiber will then be, with e= -2.07", 


29000 -193500 -193500 w -2.07 

<W =-^X5 + 19.45 X 2.38 + .844 X 19.45 X 2.38 X 2.38 - 2.07 


= 1490-4180 + 33100=30410 lbs. per sq.in. 

As above mentioned, transverse tests made on rectangular rods cut from the specimens of 
Fig. 116, gave a tensile stress in the outer fiber of 33300 lbs. per sq.in. To derive from this 
the strength of the material under pure tension, we may use the following formula due to Bach. 4 


Kb = Kz.Uq 



(9) 


in which Kb = tensile strength per sq.in. in outer fiber; 

K z = tensile strength, pure tension; 

/i 0 = a factor depending upon shape of the cross-section and whether the surfaces retain 
the cast skin or not. In this case the skin is retained and may be taken = 1.0; 

e, = maximum fiber distance; 

2 0 = distance from neutral axis of the entire section to the neutral axis of that part of 
the cross-section on one side of the first mentioned axis and tow ard the fiber 
under greatest stress. 

For a rectangular section z^ = \e x , so that we may write for the tensile stress of the rods 
which gave Kb = 33 300 lbs. under transverse stress 

Kb = 33 300 = KzV-^- = 1 AUKz, 
qqq no 

from which ^ 2= L4lT = 23500 lbS * *** ^ 


1 Elasticitat and Festigkeit, eq. (8) or Maschinen Elemente, eq. (103). 

2 The minus sign is used because the moment tends to decrease the curvature. 

3 For graphical solution, see Zeitschrift d. V. D. I. 1901, p. 164. 

‘Bach, Elasticitat and Festigkeit, art. 22, eq. (1), or Maschinen Elemente, 8 ed., eq. (137), p. 43. 











120 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Now to refer this to a cross-section like that of Fig. 117, we have = 2.07", z 0 = 1.32", and 
A z = 23500, hence for this section 


Kb = 23500^-^— = 23500X1.26 = 29700 lbs. per sq.in. 

1 . O*- 


This figure agrees very well with 30400 lbs. per sq.in., above determined by computation, 
and this would show that eq. (8) is the most reliable for similar cases. 


It should be noted that the stress will be the smaller the greater the value of r, 
the radius of the fillet. The radius of curvature should, therefore, be made as large 
as possible, which finally leads to the use of cylindrical seats for the main bearings. 

If the much simpler equations (7), which apply to straight prismatic rods, are used 
for approximate computations, the allowable fiber stress should be assumed only 
one-third as great as would ordinarily be considered the limit. If the stress in the 
dangerous section is not to exceed, say, 4200 lbs. per sq.in., the dimensions must be 
chosen such that equations (7) do not show a total stress exceeding 1400 lbs. per sq.in. 
It is evident, however, that in important cases an accurate determination of the stress 
conditions, according to eq. (8), is absolutely necessary. 

The bolts or studs connecting the cylinder to the frame must be computed with P z 
as the maximum total load. With reference to anchor bolts, see pp. 90 and 93. 


II. Cylinders and Jackets 

1. Cylinders. Material: Hard, close-grained cast iron. Allowable stress fairly high 
to reduce the thickness of walls with reference to expansion by heat. Sufficient 

For cast iron with fairly clearly defined stress 
conditions, the allowable fiber stress may be /£*== 
4250 lbs. per sq.in., in other cases less. 

Cylinders and jackets, and sometimes frames, of 
small machines are very often cast in one piece to 
cheapen the cost of construction; the larger sizes, 
commencing, say, with 8" diameter of cylinder, are 
mostly made with a separate cylinder liner, if special 
conditions, as for instance radial inlet and outlet 
parts, etc., .do not compel the use of the one-piece 
construction. A separate cylinder liner or barrel has 
the advantages that a specially suitable grade of 
iron can be employed, that the jacket and frame are 
cast witn greater ease, and that the wear on the cylinder can be taken care of more 
readily. To this should be added the special advantage, important in gas-engine 
construction, that the separate liner can expand axially independent of the jacket, 
which is very important on account of the great temperature difference in these two 
parts. To allow of this free expansion, the liner is to be rigidly fastened only at its 
inner end, the other end sliding axially in its seat. Of course, the one-piece cylinder also 
has its advantages, in that it is a stiffer construction than the other, and for that reason 
its use would seem to recommend itself especially for large engines; but greater difficulty 
of manufacture and repair, unequal expansion under heat, etc., counterbalance this gain. 

The temperature differences existing between the inner and outer walls of the 
cylinder and their influence upon the expansion or stress conditions in these important 
parts have been investigated but little. They are, however, certainly greater than is 


stiffness should be maintained. 



FiG. 118.—Cylinder of a Korting 2-cycle 
Engine. 
















































GENERAL ENGINE PART& 


121 


commonly assumed. Of all the direct temperature measurements undertaken, those 
made by Ernest Korting, in 1901, on one of his large double-acting 2-cycle machines 
are of greatest value to the designer. The cylinder of a 400 H.P. engine of this type 
was fitted with thermometers a-d, as indicated in Fig. 118. The bulbs of these 
thermometers were placed exactly at the center of the inner wall. Tubes passing 
through the jacket space and having their lower ends filled with mercury, were 
inserted as shown, so that the centers of the thermometer bulbs were from the 
bore of the cylinder. 

Table 9 


TEMPERATURES OF THE INNER CYLINDER WALL OF KORTING ENGINES 


Hour 

3:30 

4:00 

4:15 

4:30 

4:45 

5:00 

5:15 

Load on engine. 

Temperature of cooling water, outlet ° F. 

r b, 0 F. 

Temperature of cylinder wall at point . . j c, ° F. 

i 

4 

91 

167 

134 

314 

i 

4 

95 

167 

134 

316 

£ 

97 

181 

147 

321 

5 . 

8 

99 

194 

150 

330 

a 

99 

190 

143 

338 

l 

99 

198 

145 

338 

full 

101 

201 

145 

338 


The measurements were made with different loads on the engine. From Table 9 it 
is seen how quickly the temperature changes when the load on the engine varies, i.e., 
when the quantity of heat developed changes. It will also be seen that the tempera¬ 
ture of the cylinder wall near the combustion chamber is considerably higher than at a 
point about mid-stroke, in spite of the fact that the point b, on account of its position 
near the dead center, is covered by the water-cooled piston, at least for a time 
equivalent to one-quarter of the stroke. The conclusion from this is that the tempera¬ 
ture of the combustion chamber, which is subject to much higher gas temperatures, and 
which is at no time cooled from the inside, must be much higher than the temperatures 
at b The highest temperature determined at full load is 338° It is unquestionable that 
the temperature of the inner wall is higher than this, especially if the piston be not water 
cooled, or shorter than in this type of machine. If the temperature of the jacket 
wall is assumed equal to the mean temperature of the jacket water, i^.=84° F., and 

the mean temperature of the cylinder wall equal to — ^ ’ neglecting the 


temperature near the exhaust ports, the expansion of the cylinder, relative to that of 
the jacket wall will be = (173-84) X.00000593 = .000528. Even if this entire expansion 

were to be taken up by the jacket wall alone, the stress induced would still be below 
the elastic limit of cast iron, the extension of which m tension is on an average not 
below .00075 at the elastic limit. The case; however, becomes a little more serious 
when the mean wall temperature t m , including the exhaust port temperatures is used 
in the computations. U will be = (201 + 145 + 338) ^-3 = 228° F and the expansion 
relative to the jacket wall will be - (228-84) X.00000593 = .000854. This already 
exceeds the average extension of cast iron under tension by 13%. Now / ls ’ °* 
course, a fact, that in one-piece cylinders the axial stresses are not earned by the 
cylinder wall alone, but by the cylinder and jacket walls combined. This, under te 
most favorable circumstances, reduces the stresses by one-ha . n e o > 

the elastic limit strength of cast iron is considerably reduced by repeated heating a 
cooling, such as exists in gas engines, i.e., the iron grows more bntt e and is more 
liable to fracture. In general, too, the outlet temperature of the jacket water is much 


























122 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


higher than that used in the above tests, and the internal cooling of the cylinder wall 
is much less effective with pistons not water cooled. Finally, the stresses to which 
the cylinder is subjected by the explosion pressure in both radial and axial directions 
enter the problem. Under such conditions it may happen—and practical experience 
proves it—that the sum total of the stresses due to expansion by heat, and to the 
working load on the cylinder may far exceed the elastic limit strength of the material, 
leading to serious damage by fracture or warping, if the cylinder is cast in one piece. 

It is not easy to relieve the jacket walls of double-acting cylinders, with valve cages 
cast in one piece with the cylinder, of longitudinal temperature stresses. The use of a 
separate cylinder barrel is in the case of such cylinders no longer practicable, and it becomes 
necessary to divide the jacket wall into two parts, longitudinally, so that the two half 
cylinders so formed, surround the cylinder wall proper in saddle shape. The construction 
is shown in the double-acting Deutz engine, Plates IX and X. Langen and Wolf, in 
their new machines designed by Ebbs, use the method of relieving the jacket wall de¬ 
scribed in Part III under III, 1, in which case the one-piece construction maybe retained. 

Injurious temperature stresses may also be caused by injudicious distribution of 
metal, for which reason it is well, in the case of large cylinders, to design with an eye 
to symmetry, at least, as far as the combustion chamber and the adjacent parts are 
concerned. Any violation of this rule also introduces trouble owing to the fact that 
unequal distribution of metal causes irregular radial expansion of the cylinder bore, 
which of course, leads to cutting and leaking between cylinder wall and piston. 

The water jacket should, If possible, cover the entire 
Designs of Cylinders. length of the stroke, even in single-acting machines. If 

the open end of the cylinder is left unjacketed, there will 
be unequal expansion in the cylinder bore, and annoying 
vapors due to burning lubricating oil are apt to develop. 
All water spaces should be of ample width, and in the case 
of large cylinders sufficient cleaning openings should be 
provided to rid the water space of core sand in the course 
of manufacture, and of any sediment or scale that may 
form during operation. Naturally, the cold water should 
be admitted to the jacket at the hottest part of the cylin¬ 
der, and after this it should be so conducted, by means of 
baffles, if necessary, that nowhere in the jacket space shall 
there be a chance of impeded circulation or formation of 
air pockets. 


Fios. It9 and 12a-Cylinder and Cylin- Fig. 121.—Cylinder for a 60 H.P. Engine, Langen & Wolf, 

der Head for a 20 H.P. Banki Engine. Vienna. 

(The broken lines a and b indicate the (Cylinder head in one piece with liner.) 

upper and lower ends of the piston in its 
highest position.) 


































































GENERAL ENGINE PARTS 


123 



Figs. 122-125.—Cylinder for a 12-15 H P. Hornsby-Akroyd Engine of German Make. 
(Projection b is to prevent the piston from going too far back when assembling.) 


Constructive Details. The volume of the combustion chamber may be computed 
from eq. (55) p. 32. From the compression ratio e, required for the given compres¬ 
sion pressure p c , the length h of the combustion chamber, when S is the length of 
the stroke, will be 


assuming the combustion chamber to have the same cross-section as the cylinder bore. 

The explosion pressure p z , which determines the cylinder dimensions, causes tensile 
stresses in the cylinder walls, both in a longitudinal and a radial direction. To these 
should be added, under circumstances, secondary stresses due to drawing up of the 
flange nuts, the thrust of the connecting rod, unequal temperature stresses, etc. In 
the case of one-piece cylinders, the tensile stress in the longitudinal direction may be 
neglected, if the inner cylinder wall is properly designed to withstand the radial stress 
due to p z , because this latter stress is always more than twice the former. If, on the 
other hand, jacket wall and cylinder wall are separate, the jacket wall alone is 
usually called upon to carry the longitudinal stress or the stress in the axial direction, 
and since the jacket wall is usually computed on a basis of load much less than that 
due to p z , care should be taken in this construction to see that the axial stress does not 
become dangerous. This is especially important for places in the jacket wall that are 
expected to carry an additional external load or that have been weakened in any way. 

According to Bach, 1 if we let 

R a = external radius of cylinder in inches, 

Ri = internal radius of cylinder in inches, 

K z = allowable tensile stress for the material, 
and pi = internal pressure pounds per sq.in., 


1 Bach, Maschinen Elemente, 8 ed., p. 36. 



































































124 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 
we shall have the relation 

i 1+.4A • 

Ra = Ri I -—— = RtyjT— —inches .(11) 

\l_1.3 PL y K z -l.S Pi 
K z 


From this, if s is the thickness of the cylinder wall, D = diameter of cylinder in inches, and 


Pi = Pz- 


s 



inches, 


( 12 ) 


Since p z acts only at the inner end of the cylinder barrel which is nearly always 
reinforced by a flange, the value of K z may be taken correspondingly high, from 
3200-3500 lbs. per sq.in. for cast iron. 

The normal indicator diagram then calls for a thickness of inner cylinder wall of 
s = ^(Vf500 ^ 460 ~ 0 1,197 “ = .° 47 D inches.(12 n) 


To this should be added, as an allowance for reboring, from ^ to \ in. For practical 
reasons (possible shifting of cores, etc.), it is also necessary that, in the case of 
stationary machines, up to say 8" diameter, s should be at least ^ .7", for one-piece 
cylinders, and at least ^ .65" for separate cylinder barrels. 

In cylinders over 25" diameter, the thickness of the wall may, for reasons of 
economy and decrease in weight, be decreased gradually or by steps until at the open 
end 

$o = .02 D + .5 in.. . (12a) 

provided additional stresses or other circumstances do not forbid this. 

Due to the pressure p z , the inner fibers of the wall are under a greater stress than the 
outer fibers. This inequality of stress increases materially with an increase in the 

thickness of the wall. For that reason s should not be 
made thicker than necessary. Excessive thickness of 
cylinder wall is also wrong from the standpoint that the 
cooling water is less effective, consequently the conduc¬ 
tion of heat is less rapid, and troublesome overheating 
of the material may ensue. 

In large or long-stroke engines, the thrust of the con¬ 
necting rod should receive careful attention in dimension¬ 
ing the material of the cylinder. In case the dimensions 
of the Inner wall, as determined by eq. (12) should furnish a section modulus too 
small to take care of the normal pressure N due to the connecting rod, it is better to 
increase the stiffness of the cylinder wall by suitable supports or webbing than to 
increase the thickness s. 

With the favorable assumption that N is distributed over the bearing surface of 
length l , Fig. 126, the maximum bending moment will be 



Mb = Al [cV~)^W K b . 


. . (13) 

































GENERAL ENGINE PARTS 125 

In eq. (13) the maximum pressure at the supports is 

pounds.(14) 


In case a separate cross-head is used, l is the length of the shoe, with trunk pistons, 
l is the length of the piston body in actual contact, i.e., excluding any part made 
smaller to give play for expansion. 

The stiffness of the cylinder liner, or the cylinder, should be such that N cannot 
cause any noticeable bending or deformation. In order to be sure of this point it is 
advisable also to make a check computation, assuming that the force N, instead of 
being distributed over the length l, is concentrated at the middle of this length. In 
that case 

M b = N— in.-lbs.(15) 

a 


and assuming that Kb is the allowable stress, the section modulus required will be 


w =n™l >£, 

ah.b Kb 


(16) 


Assuming that the moment of inertia of the cross-section of the cylinder is I and that 
the coefficient of extension is a, (a = reciprocal of modulus of elasticity), the deflection 

under N, which will be a maximum when 6 = &i=|-, will in general be 


Nb 2 bi 2 a 

Sal 


inches, 


(17) 


If on the other hand, b<b i, 
the left support such that 


the maximum deflection will occur at a distance x from 


x = + inches, 

o obi 


(18) 


Examples. 1. A cylinder liner, with D = 25.6", for an explosion pressure p z = 356 lbs., should, 
according to eq. (12n), have a wall thickness of 

s = .047 D = .047 X 25.6 = 1.2 + f (for re-boring) = 1 y*/'. 

With reference to Fig. 126, let a = 50.3", 6 = 23.6", 6,=26.7". Then the maximum bending 
moment, due to N = = iVX280X25.6!=18400 lbs., will be 

1840 0X ^3.6 X ^^230000 in.-lbs. 

Mb 50.3 

Now the section modulus is 

W — .8 Drn^Sf 


in which D m = mean diameter-25.6 + 1^=27.2". Hence 











126 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

The maximum bending stress therefore is 


<76 = 


230000 

923 


= 250 lbs. per sq.in. 


The cylinder liner proportioned according to eq. (12w) seems therefore of ample strength 
to take care of N u but the maximum deflection / should next be determined. 

The moment of inertia of the cross-section is 


/ = .05 (D 0 4 -D 4 ) = .05[(25.6+2X1 T %) 4 -25.6 4 ] = 12 450 in. 


in which D 0 = outside diameter of the liner. 

Taking the extension coefficient of cast iron as a = — qqq'qqq the maximum deflection 
according to eq. (17) will be 


18400X23.6* X 26.7* 

3 X50.3X12450X11000000 


= .00036". 


Since the load N is not concentrated but distributed over the length l, which makes f still 
smaller, this result shows that this cylinder liner is safe without any further stiffening, provided 
the two points of support, Ai and A r , are not subject to deflection. 

The flange or collar of the cylinder liner is very often subject to stresses, induced 
by the drawing up of the cylinder head studs, which tend to break the flange along a 
section x-y, Fig. 127. The pressure due to this drawing up 
must in all cases be greater than the maximum pressure on the 
cover, which is equal to .785 Djp z , in order to retain at the 
moment of explosion a sufficiently high pressure Pf on the pack¬ 
ing surfaces to prevent them from leaking. Just how much the 
total pressure under the cover nuts should exceed the axial in¬ 
ternal cover pressure for the purpose indicated, cannot be stated 
off hand. The kind of packing, rigidity of the cover, and the 
distance from packing to stud circle, all have an influence on this 
question, so that in favorable cases an addition (£) to the internal 
cover load of 5% will be sufficient, while in very unfavorable cases 50% may not be 
enough. With good design, practice has shown that an addition of f = from .1 to .2 
to the internal cover pressure is sufficient to maintain tight joints. 

The pressure on the joint due to tightening up of bolts is then 

P a = (1 +t)pz X.785l.lp z X.785 Dj pounds.(19) 

The flange must be strong enough to resist this static load P a . The internal cover 
pressure does not enter the problem, because in cases where a separate cylinder liner is 
employed, the axial cover load is taken care of by the jacket wall. 

The force thus developed, induces stresses in the cross-section x-y as follows: 



Fig. 127.—Dimensions 
in Inches. 


A bending moment due to P a and the arm z\ 

A tensile stress due to the component P n = P a sin a, and 
A shearing stress due to the component P s = Pa cos a. 















GENERAL ENGINES PART 


127 


The shearing force P 8 should not be neglected, because the arm z of the force P a 
is usually small in comparison with the height h of the cross-section x-y. The various 
stresses are as follows: 

From the bending moment Mb = P a z and the section modulus W = 0 f th e 

section, we have the maximum bending stress 


M b 

b= W 


6 PgZ 
ti Dob? 


lbs. per sq.in, 


( 20 ) 


From the normal force P n and area of dangerous section x-y equal to f=7zD 0 h, the 
tensile stress is 


Pn 



P a sin a 

tzDqIi 


lbs. per sq.in., 


( 21 ) 


and from the shearing force P s together with area /, the shearing stress is 


r= 


Ps 

f 


Pa cos a 
7lD 0 h 


lbs. per sq.in, 


( 22 ) 


Combining first, <jfc+<r z =< j lt and then this with r according to Bach’s equation 
(see p. 168, Table 15) gives finally the resultant stress, 


eyes. = -35 <ti + .65V <7i 2 + 4(« 0 t) 2 .(23) 

This must not exceed the allowable bending stress Kb. For cast iron, for which the 
allowable tensile and shearing stresses may both be taken equal to A n = A s = 2800 lbs. 
per sq.in., the value of the factor <*0 is 

K n 2800 _ 

a '° 1.3A S 1.3 X 2800 

Example. 2. The cylinder liner computed in Example 1 has the flange sketched in Fig. 127. 
With the dimensions there given and an explosion pressure p z assumed at 356 lbs. per sq.in., we 
shall have the following: 

P a = 1.1X 356 X .785 X 27.2 1 = 220 000 lbs.; 

Mb = 220000X.79 = 174000 in.-lbs.; 

Pn = 220000 X sin 20° = 220000 X .342 = 75 300 lbs.; 

P s = 220 000 X cos 20° = 220 000 X .94 = 207 000 lbs.; 

• 89 2 X1 89 2 

/=89.2X1.89 = 169 sq.in.; W = — : —-= 53.3; 

174000 75300 ..... 

Ob= ~ -= 3260 lbs. per sq.m.; o z =——=446 lbs. per sq.in.; 

53.3 low 

t=^ 07^29 = 1225 lbs. per sq.in.; <76+<7*=< 7 , = 3706 lbs. per sq.in. 

169 

Hence combined stress is, from eq. (23), 

< 7 res . = .35X3706 + .65V3706 2 + 4(.77 X1225) 2 = 4000 lbs. per sq.in. 













128 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

For the determination of the stress in the dangerous sections x-y, of the flanges 
of cylinders under internal pressure, Professor Bach 1 gives the following general 
equation: 

3 J^SE>K b .(24) 

hi ~ 

Here p is a correction coefficient which Bach determined as .8 on tests of similar cast-iron 

flanges, Rb = ^r-, p = Pz, and z and hb are according to Fig. 127. 

& 


If this equation had been used in Example 2, the result would have been 
3X.8X14.2X356X.79 


°res. - 


(1.73) 2 


- = 3200 lbs. per sq.in. 


The first method of computation therefore gives a somewhat greater factor of safety, which, in 
this important place, is quite desirable. 


If the flange is cored out for water passages, the amount of cutting away must be 
taken into account by subtracting it from the circumference tcD 0 of the gravity circle. 
The height h under such circumstances, is usually increased so that the flange wall 
separating any two water spaces becomes a radial rib or web which materially assists 
temperature equalization. (See Figs. 119 and 120, p. 122.) 

2. The Jacket Wall is in the direction of its axis under a tensile stress 

Pz = pzX.785 D 2 lbs. 


Its full cross-sectional area is expressed by (see Figs. 128 or 129) 

/ = .785(D a 2 — Di 2 ) =xD m s' sq.in. 

The tensile stress therefore is 

p z X.785 D 2 p z D 2 

02 .785 (Da 2 — A 2 ) Da 2 -A 2 ’ * 

PzX. 785 D 2 p z D 2 

0r> ~~ ~KDm7 ~ 4A*? . 

From (25a) 

v-J? !? 8 .. 

4 DmKz . 


(25) 
(25a) 

(26) 


If K z = allowable tensile stress ^1800 lbs. per sq.in., and p z is assumed = 356 lbs. per 
sq.in., the thickness cf the jacket wall will be 


, ^ 356D 2 

“ 4 X 1800 Dm 


D 2 

20 Dm 


(26 n) 


In practice s' should not be less than T \" for small engines nor less than for 
the larger sizes. The cross-sections of the jacket wall weakened by cleaning openings 
or otherwise should be rechecked for strength. In the case of one-piece cylinders the 
tightness of the cylinder wall proper is often tested by hydraulic pressure. If the test 
pressure employed is equal to say 1.5/> ? = 650 lbs., and the allowable bending stress is 


Bach, Maschinen-Elemente, 8 ed., p. 691. 















GENERAL ENGINE PARTS 


129 


taken at 9250 lbs. per sq.in. for a single application of such test, the jacket wall 
thickness necessary to satisfy these requirements will, from eq. (12), p. 124, be 


s' 25 D a in. 


(26a) 


The thrust of the connecting-rod tends to bend the jacket wall as it does the 
cylinder wall, and the stresses and deflections involved can be determined in the 
manner already shown. For cylinders which are entirely supported in the frame, the 
axial tensile force P z (see pp. 103 and 113), together with the practical requirements 
of the foundry, will call for a thickness of jacket wall sufficient to take care of N. 

A closer investigation of the influence of N must, however, be made in the case of 

large and long stroke engines having overhung cylinders, because in this case one end 
of the jacket wall must take up the entire bending moment. The dangerous section is 
then close to the connecting flange, and in the case of weak frames 
it may even be found in the flange seat of the frame itself. 

The flange used to connect the cylinder, or the jacket wall, with 
the frame has a metal to metal seat and is not called upon to resist 
gas pressure against leakage. Its dimensions therefore depend entirely 
upon the axial cover pressure P 2 = p*X.785 D 2 , if the normal pressure 
N does not enter the problem. The stresses in the dangerous section 
of this flange may be found in the same manner as for the cylinder 

barrel flange, pp. 126 . and 127. Bach’s equation (24) there given may be written 

(see Fig. 128) 

fiPaz' HJ X2xRbh\ 2 Kb, 



so that the maximum bending stress will be 

3 fiPaZ' _ fiPgZ' 


Ob = 


TiRbhi 2 1.047Rbhi 


o - K b . 


(27) 


The bending moment P a z' tends to deform the flange, as shown in Fig. 128, and 
it may happen that the inner cylinder or liner is thereby deformed in the bore. 
Sufficient rigidity is therefore of great importance. This may be obtained either by 
increasing the thickness, s, of the cylinder wall under the flange seat, by using radial 

webs, or by using heavy bosses in the flange for the bolt 
holes. All such stiffening should, however, be left out of 
account in strength computations. 

The bending moment in the flange for the cylinder cover 
or head will be a minimum when the bolt circle coincides 
with the gravity circle of the cross-section of the jacket wall. 
In that case the axial load acts as a tensile stress only, 
and w r hen two-piece cylinders are used only that part of 
the jacket flange lying within the bolt circle is under a 
bending moment due to the pressure Pf necessary to keep 
a tight joint. At the moment of explosion, however, this 
pressure Pf is relieved by the load on the cover = p 2 X.785 D^, 
so that at this instant the stress remaining is only that due to the necessary excess 
of P/ over the internal load on the cover. 

The stresses are much more severe when the diameter of the bolt circle is greater 
than the outer diameter of the jacket wall. The flange is then under a heavy load 
due to the drawing up of the cover studs, which load induces bending, shearing, and 


Dr-27- 

f 

<?-4J4 - J 

*7' 4 -— 1 

-rhr -— 

rr 

•‘1.27'f i 

. — 


i * 


q-35.3 



M 1 

k 



D25.6- 


Ho 




ST'IP 


Fig. 129.—Dimensions in 
Inches. 
























130 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


tensile stresses, as was shown from the liner flange on p. 126. The following example, 
in connection with Fig. 129, explains the conditions then existing in detail. 


Example. The cylinder liner, Fig. 126, computed in the previous example, is seated in the 
jacket flange, at the head end, as shown in Fig. 129. The pressure P a required on the joint 
was on p. 127 determined at 220000 lbs. From this we can derive at once 

P n =220000 cos 30° = 220000X .866 = 190500 lbs.; 

Pd=Pa = 220000 lbs.; 

P s = 220 000 sin 30° = 220 000 X .5 = 110 000 lbs. 


The area of the dangerous section x-y =/= 33.3 X * X 2.44 = 255 sq.in., and its 
modulus is 


W = i X 33.3 X * X 2.44 2 = 104.0. 
Bending moment Mb = PdZ = 220000X 1.27=282000 in.-lbs.; 


r = 


Ob = 


110000 

255 

282000 

104.0 


=430 lbs. per sq.in.; 
= 2700 lbs. per sq.in.; 


190500 menn 

o 2 = -— = 7o0 lbs. per sq.m.; 

Zoo 


<r,=u 2 +<16 = 750+2700 = 3450 lbs. per sq.in. 


section 


Hence the combined stress is 


a re s. = .35X3450 + .65V'3450 2 + 4(.77 X 430) 2 = 3590 lbs. per sq.in. 


In this case, therefore, P s or r could have been neglected, 
r 

The cross-section x-y', should next be checked. The shearing stress for this section is 
found to be 


220000 

T “300rX1.57" 1440 lbs - P 61, "» jn - 
The combined stress in x-y is close to the allowable limit. 


In one-piece cylinders it is often possible to make the radius of the cover' stud 
circle equal to the mean radius of the jacket space, or to make it so that the studs 
are close to the inner jacket wall. Under these conditions the bending moment of the 
flange is very small. An arrangement of this kind combined with sufficient rigidity 
for the frame flange can, however, only be had in the smaller sizes of engine. 

Where special lightness of machine is desired the jacket is sometimes made out 
of sheet steel fitted over the head and crank-end flanges of the cylinder. This 
construction is not easily damaged by freezing of the cooling water, and it also admits 
of thorough cleaning of the jacket space from scale and mud. It has the disadvantage, 
however, that the steel, especially at the ends, is apt to rust through quickly, a 
process which is only temporarily arrested by painting or tinning. 

The connecting bolts between frame and cylinder may be of wrought iron, if well 
made. The allowable stress may be taken at 5000 lbs. per sq.in. of cross-section, as 
measured at the bottom of the thread. It should be remembered in this connection 
that a few strong bolts are better than a larger number of correspondingly smaller 
tensile strength. The smallest number of bolts should be four for small engines. 
English designers, even for the larger sizes, often use only two bolts, which are placed 
in the horizontal bisecting plane, that is, in a position not correct with reference to 
connecting-rod thrust. In general, the less the apparent rigidity offered by the design 
of the cylinder and frame, the greater should be the number of connecting bolts used. 

For cylinder-head packings and cover studs, see pp. 137 and 138. 







GENERAL ENGINE PARTS 


131 


III. Cylinder Covers (Cylinder Heads) and Stuffing Boxes 

Material is usually tough, close-grained cast iron. During the first years of the 
development frequent cylinder-head fractures drove the designer to the use of cast 
steel. Since then, however, much simpler forms for the heads, which can with perfect 
safety be made of cast iron, have been developed. 

The allowable stress should, in such places where it can be determined with any 
accuracy at all, be kept moderately high, for cast iron say ZG> = 3500 lbs. per sq.in. This 
is done in order to get a construction not too rigid against expansion by heat and to obtain 
better conduction of heat through the walls. In general, however, the less the certainty 
with which the acting forces can be determined, the lower should be the value of Kb. 

The distinction usually made between the term cylinder cover and cylinder head is 
that the former has a shape approximating a flat plate, or a plate having but simple 
curvature. The latter, on the other hand, contains the valves, and its shape is there¬ 
fore much more complex. In fact in some cases the dimensions of these heads in the 
axial direction is so great that they practically contain the combustion chamber. The 
triple duty (cover, valve cage, and combustion chamber) so imposed makes the 
cylinder head a machine member of the greatest importance and subject to the most 
severe stresses. Its design, for machines above a limited size, is further complicated 
by the fact that it is difficult to apply an even fairly rigid method of computation, 
and everything depends upon feel and judgment of the constructor. 

Simple cylinder covers are obtained when the valves are placed at the side of the 
cylinder, leading into the combustion chamber. This method is, however, only possible 
in the smaller sizes and where moderate compression is employed. In large machines 
with high compression there is not enough room for the valve ports in the circumferential 
surface of the compression space. Besides this, these ports weaken the cylinder, 
prevent the use of separate cylinder liners or bushings, and form pockets in which 
exhaust gases are apt to remain and in which ignition is imperfect. Not quite so bad 
in this sense is the new arrangement of placing the valves at top and bottom at the 
end of the cylinders. But even this scheme is open to some objections, as, for 
instance, the breaking up of the combustion chamber. It is also to be feared that, although 
by this design the cylinder heads are relieved and made much simpler in construction 
and safer in operation, the trouble has only been transferred to the cylinder, itself a 
machine member which already requires the most careful consideration in its design. 

The design of the cylinder head is in the main subject to the following three 
fundamental requirements. 

(а) Thermodynamically proper form of the combustion chamber, so that the distances 
for flame propagation in all directions from the igniter shall be the smallest possible, 
that there shall be no dead spaces or corners and inadequately cooled projections, and 
so that the movement of the gases from the combustion chamber into the cylinder 
proper shall occur in an axial direction as closely as possible. The inner form of the 
compression chamber should therefore approximate a hemisphere, or a short cylinder 
(length = diameter). The inner walls should be unbroken and smooth, and care should be 
taken not to deflect the stream of gas laterally as it enters the cylinder bore (see Part V). 

(б) Correct arrangement of the openings for valves and igniters. The unavoidable 
valve ports should reduce the enveloping surface of the chamber as little as possible. 
With due reference to practical requirements, these ports should be so distributed that 
there shall be an approximately symmetrical distribution of metal in the central 
cross-sections, so that the water-cooling may be equally effective on all sides. Local 


132 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


accumulations of heat causing unequal expansions are thereby. prevented and there 
will be no pockets out of the path of the jacket water. The jacket wall, in proper 
places, should be provided with sufficiently large cleaning openings, through which the 
space may be cleaned very thoroughly of core sand, and any deposit by the cooling 
water in the way of scale or slush can be removed from time to time. (Failure to 
remove such accumulations from the cooling surface has already resulted in serious 
accidents in operation.) 

(c) A general shape of the cylinder head such that a little flexibility prevents the 
occurrence of high stresses. The design should be as far as possible of regular form, 
and of sufficient strength and stiffness, but the inner and outer walls should not be 
so rigidly connected that the latter cannot accommodate itself to the greater expansion 
by heat of the former. (Compare p. 120.) Unnecessary piling up of metal and 
abrupt changes in the thickness of cross-section should therefore be carefully avoided, 
the two walls should be connected only where necessity demands it, and rigidity 
should be provided for only in places where it is required (flanges or valve seats). 

The third requirement above outlined is, of all things connected with gas-engine 
design, the most difficult to properly carry out. The designer is thrown entirely upon 
judgment and trial, and even the most experienced are not safe against failure. The 
construction of a large cylinder head, more than that of any other machine member, 
therefore calls for close co-operation of drafting room and foundry, but even then the 
most careful work in the casting of the first few heads of new design does not always 
guarantee immunity against fracture in operation. 1 

Designs of Cylinder Covers and Heads. (See also Figs. 71, 124.) 



1 Valuable information from practice is contained in a lecture by Reinhardt, published in Stahl 
and Eisen, 1902, No. 21. 












































































































/3.2 










































































































































































































































































































































































134 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 



Figs. 144-147.—Cylinder Head for a 300 H.P. Gas Engine built by L. Soest & Co., Reisholz, Diisseldorf. 

Material is cast iron. (The Frame for this engine is shown in Figs. 87-89, p. 107). 

Constructive Details. Strength computations give useful results only in the case 
of the simpler covers. As regards cylinder heads no close mathematical treatment 
is possible on account of complexity of form and uncertainty of casting stresses. The 
designer is forced to rely upon experienced judgment and a few tried practical rules. 

The cylinder cover of simplest form is a plain circular plate, which may be 
considered as a beam under transverse stress, supported at the ends of a diameter. 
From what has preceded (p. 126), the cover-nuts must be drawn up tightly enough, 
so that at the moment of greatest pressure p z enough pressure still remains on the 
joint to keep it tight. If, as before, we assume the total stress in all of the studs 
to be 

Pd= (1 +OP* X.785 Dylbs., 
then each one of i studs has to carry a load of 

p _Pd_ (1 + Oy>sX.785-D/ lbg 


i 


(28) 










































































































GENERAL ENGINE PARTS 


135 


The pressure P d produces a constant bending moment 


M b = 


Pde 


iPe 

TZ 


in.-lbs., 


which leads to the requirement that 

WKb >^ = &. 

TZ K 


(29) 


(30) 


In these equations e is the radial distance from center line of the stud to the 
middle of the packed surface (see Fig. 129, p. 129), and fi a correction coefficient 
which, according to Bach, 1 may be made-f for cast iron. If the cross-section of the 
cover is of simple rectangular shape, without water jacket, whose dimensions are, 
thickness = h, and diameter (length) =D d =2R d , the section modulus of the central 
section will be 

W=tDJi 2 =iRJi 2 . 


We may then write 


KtXiRha 2 ^ 


j uPtie _ idPf 


and consequently the bending stress in the section, independent of the. explosion 
pressure, p z and depending only upon the stress P d in the studs is, putting P = f, 


3.6 Pae 3.6 iPe 
° b ~ R d h 2 7t - Rdh 2 7T 


lbs. per sq.in 


(31) 


At the moment of explosion the internal load on the cover is p z X.785Dj lbs., which 
tends to bend the cover outward, and thus decreases the pressure P f upon the packing 
to a remainder tP f = iP 0 , where f<l represents the percentage of excess pressure as 
before explained (p. 129). The maximum total load on the cover is therefore 

P^=(l+f)p z X.785D^ = 'iPo + P z X.785D^ lbs.(32) 

Of this amount the force acts along the semi-circumference * §'=«»/; while the 

l oac l -7 85 D /P* is uniformly distributed over the semi-circular area For the purpose 

of simplification, however, we will assume that both forces act at the center of 
gravity of the semi-circle, that is, at an arm a = if-.4244B/ from the center of the 
cover. The maximum bending moment then is 

Mi , (l+Op.X-785 Dj x i2U R/ in ,i ha . (33) 

or, since D t =2R„ so that .5D}=2Rj, and 4244X.7S5 = i, we may write, 

jr ' (1 +0p.X2 h'fh'f _ '1 , p3; n _ lha .(33a) 

3 3 


1 Bach, Maschinen-Elemente, 8 ed., p. 37. 













136 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 
But it has been shown above that the section modulus of a plane cover is 


W=\Rdh 2 . 


Hence the bending stress resulting from Mb' will finally be 


2(1 + 1) p z Rf 

° b ~ 


lbs. per sq.in, 


(34) 


Bach, 1 in his computations in this case, proceeds as follows: 

Combining the moment of the internal pressure 

= .5*Kh>,xf —=!R/P*, 

O TZ o 

the moment of the pressure remaining on the packing at the instant of explosion 

= 5iPo mjisR, 

TZ TZ 

and the moment of the total stress iPo+7zRfp z in the studs 

= .5 (iPo+xRjp,)—, 

TZ 

the resulting bending moment will be 

M b '-(iPo + rRjp ,)—-f R?p,-t'P 0 ^ 

71 6 TZ 

=RjR, P ,~~Rj Pl +iP 0 ^ .(35) 

If the distance e, see Fig. 129, from stud circle to middle of packing is small, 
the pressure tPf=iPo will also be small. This is to be aimed at if possible. If this 
pressure is neglected, and ju = $ is introduced into eq. (35), this may be written 

Mb' = $X.RfRdPz — %R/p z ^WKb .. (36) 


The maximum bending stress for the rectangular cross-sections is then 

„ [ JRjR-ipz-^RfPz 

\Rdih 2 . 


(37) 


In order to simplify eq. (37), Bach substituted for Rf (the radius up to the middle of 
the packing), the stud circle radius R,i = R/ 4- e. Then 

Mb=%RdPz in.-lbs. ... (38) 


1 Bach, Maschinen-EIemente, 8 ed., p. 683. 










GENERAL ENGINE PARTS 


137 


A simple flat cast-iron cover therefore requires a thickness 

***>Sfe. 

or, putting p z = 350 lbs., and Kb — 3500 lbs. per sq.in., 


(39) 


*5X3500 


= .346 Rd~\Rd ins. 


. (39n). 


Since, however, especially in case separate cylinder liners are employed, the distance e 
is apt to be considerable, it is recommended that eq. (34) or (37) be used. 

For covers which are water-jacketed, the moment of inertia of the dangerous 
section, as far as this can still be determined, must be found, and from this the 

section modulus W = - is found in the ordinary way. If the cover is pierced by only 
e 

a single and comparatively narrow axial opening, it may be assumed in the computa¬ 
tions that the connection between outer and inner walls is sufficient to overcome any 
weakening due to the opening, i.e., both walls may be considered unbroken, since 
experience has shown that a fracture never passes through the hole, but is always 
found outside of the connecting wall. 1 

But if there are in the water-cooled cover several irregularly placed chambers, or 
if there is a comparatively large central opening, computation fails and estimating has 
to be resorted to. For cylinder heads in their complicated forms this latter method 
of design is at any rate the most common. Generally reliable practical rules cannot 



Fig. 148—Methods of Packing Cylinder Heads. 


be obtained in this case even from designs successful in practice, as too many factors 
enter the problem. If the cylinder head has a free-connecting flange, which is the 
case in at least all of the larger machines, special attention should be paid to the 
transition from this flange to the cylindrical surfaces of the head. The distance e 
above explained, see Fig. 129, may be so great that considerable bending stresses are 
set up at this point. It is usually possible to satisfactorily check this flange by itself, 
by the method outlined on p. 126. A second cross-section under severe stress is 
usually formed through the center planes of the valve seats, especially if’the valves 
are large, and arranged concentrically one above the other. The danger of fracture at 
this point is emphasized materially when the valve cages have conical form, the seats 
strongly pressed together having a tendency to cause fracture by wedge action. 

The inner surfaces of the heads should be curved if possible. This not only adds 
to the strength but is also favorable to free expansion and to combustion. Large 
internal flat suriaces are dangerous. In determining the thickness o t e wa s a y 
external load brought to bear by valve housings, valve gear, pipe connections, etc., 
should not be forgotten. 

Fig. 148 a-g shows several forms of cover packings Of these the first three can 
be used only in the smaller sizes of engines. In general, only those forms of packing 


* Compare also Bach, Maschinen-Elemente, 8 ed., p. 685. 






138 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


in which the packing material cannot be driven outward by the gas pressure offer 
sufficient safety. The simple tongue and groove arrangement, Fig. 148 d, is probably the 
most common, but costs more and is not so easy to make as the loose-packing ring, 
Fig. 148 e-g. Such a ring can first be accurately made in itself and then easily 
ground to a fit on the packing surfaces of cylinder and head. Between the surfaces of 
the ring and the flange it is usual to place a thin sheet of asbestos paper. A thick 
layer of such material, much above should not be used because the studs have 

to be drawn up too tightly to keep it from blowing out or to keep the gases from 
breaking through it. It is evident that to obtain a good joint it is necessary in the 
first place to machine these surfaces as accurately as possible, that is, to produce a 
metallic joint as far as practicable, and then to use only material of good heat- 
resisting quality, as asbestos, vulcanized fiber, etc., for the packing. 

After many years of failure, forms of stuffing-boxes have lately been developed 
which completely satisfy the requirements of present day operation. It has been 
shown that only cast iron is suitable for the packing rings, and that all of the softer 
alloys will not stand the wear. The rings are given but very little spring, and 
tightness of joint is secured by a close-fitting of the bore of the ring to the rod, and 
the use of a number of rings in series. A good method of forced lubrication, distrib¬ 
uting the oil at several places along the length of the packing, is essential to good 
operation. 

It is of advantage not to house the packing in the cylinder cover itself but to 
use a separate cage, and to provide this with separate water-cooling arrangements. 
This permits a definite control of the mean temperature of the gland. The rule should 
be departed from only when the head or cover is free from valve and igniter openings, 
thus forming a simple body with wide and unobstructed water spaces. But even then 
separate water cooling and forced lubrication should still be adhered to. 

If a special design is desired it is w’ell to use as a model tried forms of stuffing- 
boxes such as are built ready for installation by a few firms. For large gas engines 
the Schwabe stuffing-box, Fig. 149, is most commonly used. The dimensions of this 
gland necessary to fix the dimensions of the cylinder cover are given in Table 10. 

Design of Stuffing-boxes. 



Fig. 149.—Schwabe Stuffing Box for Horizontal Engine. 


























































GENERAL ENGINE PARTS 


139 


Table 10 

PRINCIPAL DIMENSIONS, IN INCHES, OF SCHWABE’S GAS-ENGINE STUFFING-BOXES 



The cylinder head-studs are made mostly of best wrought iron. In many cases 
lack of room for the nuts compels the use of a smaller diameter of stud, when soft 
steel may be employed, in which case the allowable stress, based on the cross-sectional 
.area at bottom of thread, may be taken at from 5500-6500 lbs. per sq.in. 1 

Number of studs, if cylinder head is sufficiently rigid, about 


i = .25D + 4, 


(40) 


Many works, how T ever, 'use a smaller pitch, up to 


i = .5 D+4, 


. (40a) 


D=diameter of cylinder in inches. 


To protect the packing, the pitch between studs should not exceed say 7 . Of 
course in case of necessity greater distances, up to 10", may be employed, but with 
such spacing, even rigid flanges do not always prevent the opening of the joint. 


‘Translator’s Note. The use of soft steel for such purposes is probably universal in this 
country, wrought iron being little used in machine construction. 




































140 DESIGN AND 


CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


IV. Pistons, Piston Rings and Piston Rods 


Material for the piston heads, usually a good grade of close-grained cast iron; 
for in the case of large engines, now and then, also cast steel. White metal lining on 
the bearing surfaces for both of these materials is sometimes used For piston rings 
a medium hard, but not too brittle, cast iron is best. Soft steel, with or without 

white metal covering, is sometimes employed. 

The stresses occurring can be definitely determined only m rare cases, as for in¬ 
stance for the face of the piston. For cast iron the allowable stress may be taken 


at K b = from 5000 to 6000 lbs. per sq.in. 

1 The pistons of single-acting internal-combustion engines are long, open trunk 
pistons, which, besides taking care of the gas pressure, are also called upon to 
carry the thrust due to the connecting-rod. In some of the later designs of large 
single-acting engines, the rod thrust is again transferred to a separate cross-head, and 
the bearing surfaces of the piston are often further relieved by carrying the weight of 
the piston on the cross-head and on a tail-road bearing. Quite as.sound and worthy 
of imitation is the tendency to provide all of the larger trunk pistons with water¬ 
cooling arrangements. Water cooling was for a long time looked upon as a serious 
complication, not only in operation but also in construction, and for that reason was 
not employed except in cases of necessity. But since the advantages of cooled pistons 
are more clearlv understood, they are now gradually coming into use for the smaller 
sizes of machine, down to 20" diameter. Above 150 H.P. all engines should be provided 
with water-cooled pistons. The main advantages of this type of piston are to be 
sought mainly in greater reliability of operation and in greater durability. Cooled 
pistons allow of greater compression, and by changing the outlet temperature of the 
cooling water, give a means for controlling the play between piston and cylinder wall. 
They decrease friction and consequently, also, the consumption of lubricating oil, and, 
lastlv. do away with radiation and prevent the formation of annoying oil vapors. 

Double-acting pistons are, in general, much shorter than the trunk pistons of 
single-acting machines. Their length is mainly determined by the number of necessary 
piston rings, unless, as in the Korting 2-cycle, the piston controls ports. A length 
of piston greater than that required for the rings does not offer any further advan¬ 
tages. Excessive length increases the weight of piston and its water contents, i.e., 
increases the weight of the reciprocating parts, increases the load on cylinder walls, 
piston rods, and stuffing-boxes; complicates the manufacture and increases the difficulty 
of properly lubricating the piston surface. In the building of large engines the method 
of supporting the weight of the piston on the sides of the cylinder is being gradually 
abandoned. Instead, the piston is fitted in the bore with comparatively large clearance 
and its weight is taken up by special bearings outside of the cylinder. The office of 


the piston rings is then merely to make the piston gas-tight. 

A well-designed system of stiffening webs is of special service in the case of 
double-acting pistons. This should act as an effective stiffening for the entire body 
of the piston, especially between the face and the rod boss, and should at the same 
time promote the circulation of the cooling water. The core and cleaning openings should 
be placed in the sides of the piston and not in the face, to prevent any burning in of 
fastening screws, and any weakening of the face. Cleaning openings are in any case a 
bad feature of piston design, and their use may be avoided by using a two-part piston, 


GENERAL ENGINE PARTS 


141 


see Fig. 169, p. 142. Unless the cooling water is especially impure, a troublesome deposit 
of scale is of uncommon occurrence, because the rapid movement of the water, in 
spite of high temperature, does not allow the deposition of much scale-forming material. 

The proper fastening of the piston body on the rod offers some difficulty on 
account of water inlet and exit openings encountered in the rod at this place. 
Further, the high explosion pressure necessitates rather large connecting surfaces to 
resist the axial pressure, if the unit pressure on these surfaces is to be kept within 
reasonable limits. Various methods of solving this problem are shown in Figs. 167-170. 

The fits at the ends of the piston rod for the cross-head and tail-road bearings 
should be from to f" smaller in diameter than the rest of the rod, in order to be 
able to slip on the stuffing-boxes and cylinder heads after the rod has had to be 
turned down on account of wear. 

Design of Pistons and Piston Rods. 



Figs. 151-154—Piston for a 6 H.P. Horizontal Engine. 



Figs. 155-159—Piston for a 100 H.P. Giildner Engine. 

























































































































142 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 




Fig. 162.—Water-cooled Trunk-piston, water 
inlet and outlet through the tail-rod. 



. 3 £> 



Fig. 163.—Tandem Pistons with Carrier Pistons and Cross-head, 700 H.P. Niirnberg Engine (old type)- 


Figs. 164-166.—Water-cooled Disk-piston, 700 H.P. Niirnberg Engine, old type. See also Fig. 163. 



Figs. 167 and 168.—Piston and Piston Rod, 300 H.P. Double-acting, 4-cycle Engine, built by Gasmotoren- 

fabrik Deutz. (See also Plates IX and X). 



Fig. 169.—Disk Piston for 1500 H.P. Double-acting, 4-cycle Simplex 
Engine. (Stroke 55.2", Assembly Drawing in Part III.) 

This method of piston construction has the following advantages: the 
two parts can be thoroughly cleaned of core sand during construction 
and of scale during operation, there is no thread on the rod and no nut 
in the combustion space, and lastly the holding surface on the rod is 
normal to the pressure on the piston. 




















































































































































































































GENERAL ENGINE PARTS 


143 


Constructive Details. The general formulas for the design of 
cross-heads may be applied to the design of trunk pistons. The 
length L of such pistons mainly depends upon the connecting- 
rod thrust, since this determines the unit bearing pressure K. 
We must have 


N n 


DL 


lbs., 


(41) 


in which N m&x is the maximum value of .the connecting-rod thrust. 

It can be shown, p. 158, that, for a ratio of connecting-rod to 
crank = 5, 

lbs.,.(42) 


in which P z = maximum explosion pressure. 

Hence the piston length L must be 

T pzX.785 D 2 p?X.0785 D . 

L = KrT-T0KD - K - lns - • • • (43) 


In this L may be taken to represent the real carrying length 
of piston, i.e., the total length of piston after subtracting the 
part containing the ring grooves and that part of the inner end 
usually turned down somewhat to allow for expansion. These 
amount to about one third of the total piston length. But this 
cutting down of bearing surface may also be taken into account 
by keeping the unit pressure K down to a low figure, in which 
case L from the above formula may stand for the entire length. 
This last method is the simpler of the two and will be used here. 

According to the experience gained in steam-engine practice 
with independent cross-heads, the maximum unit pressure K 
should not exceed from 35 to 45 lbs. per sq.in. In view of the 
fact that the main office of the trunk piston is to maintain a 
gas-tight packing, and that on this account the bearing surface 
must be constantly kept in the best possible shape, the pressure 
K for such pistons must be kept considerably below the above 
figures, in spite of the fact that N max occurs only every fourth 
stroke and that in general the total thrust is not serious. In 
large trunk pistons it is well not to exceed 21 lbs. per sq.in. 


For this pressure the minimum length of piston then is 

L> PzX •^ > 1 785 -— = .00374 p 2 D ins.,.(43) 

and for the normal diagram, with p z = 356 lbs., 

L> .00374 X356Z)=~ 1.3 D .(43n) 


In small engines in which longer pistons offer no special 
difficulties, L may be made as great as 2.5 D, that is, K = 



Fig. 170.—Piston and Rod for a 1200 H.P. Double-acting 4-cycle Simplex Tandem Engine. (Assembly Drawing in Part III.) 









































144 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

from 10-11.5 lbs. per sq.in. In large engines, however, excessive length of piston leads 
to the difficulties already pointed out, i.e., excessive weight of reciprocating parts, 
inequality in expansion by heat, trouble in proper fitting and in uniform lubrication 
\he weight of uncooled pistons in horizontal machines usually does not reac 
75 lbs. per sq.in. of projected piston area LD, and therefore has no appreciable 
influence on the value of K. Where water cooling is used, however it may under 
certain conditions, so greatly increase the weight of the piston that it should be take 

into account. , ... u k,. 

The thickness of the face of the piston, without webbing, may be found by 
considering it a flat round plate supported at the edge and uniformly loaded with the 
force of explosion p,. The dangerous section of such a plate is found along a 
diameter, as is the case for simple cylinder covers. This section is under the action 
of two moments acting in opposite directions. The first is due to the fluid pi ess 

act . at an arm equal t0 I* go that the moment is 
8 2 671 


M bl = i R 3 p z . 


The second results from the reaction at the supporting edge, and is equal to the 

force ^ acting at an arm of —, so that its moment is 
2 

Mtn^RZpz. 

The resulting moment is therefore 

Mb = R*pz-i R 3 Pz = 4 R 3 Pz .( 44 > 


The section modulus of a flat plate of thickness, d, b 


From the relation 


W = $Dd 2 =%Rd 2 . 

MiSWKb or ^ R 3 p z £ £ Rd 2 Kb, 


(44a) 


we can therefore derive the required thickness of piston face 

S - R '% . (43) 

Inversely, with a given value of d, the unit stress in the section will be 

ab = ^^-^Kb .( 46 ) 

i Batch modifies eq. (45), writing it d>r\p introducing a coefficient, p, whose 

value for a plate simply supported at the edge is —, and for one rigidly fastened at the 
edge is In eq. (45) we have assumed the average value /u=1.0, and have taken 
the radius of free length R = ^ (instead of r = radius of the inner surface of the piston 


1 Maschinen-elemente, 8 ed., p. 37. 








GENERAL ENGINE PARTS 


145 


face, as Bach has done). This assumption, for most conditions excessively safe, 
allows the use of a comparatively high value for the safe stress K h = 7500 lbs. per sq.m.’ 
and putting p g = 356 lbs., we obtain the special equation 

d = R ^-- 22R .( 45 «) 


In case the piston is webbed, the proper section modulus, instead of \Dd 2 is 
to be used in eq. (44a). But care should be taken not to overrate the strengthening 
effect of such webs, on account of additional temperature and internal stresses and 
the effect of the thrust N, none of which have been allowed for in the above 



Fig. 171.—Fractured Piston Head, Experimental Engine, 27.5" Cyl. Dia. 


, *4 '■ * 

discussion. That such overrating is possible is sufficiently shown in Fig. 171, which 
represents the face of the piston of a large marine engine (Loutzki), which was 
completely shattered by an explosion much more severe than the ordinary. This 
occurred when, during the trial of an engine operating on a new constant-pressure 
cycle with a maximum pressure of about 350 lbs., the supply of oil, owing to some 
irregularity, became such as to suddenly change the slow combustion into an explosion, 
during which a maximum pressure of about 1400 lbs. per sq.in. seems to have been 
generated. An occurrence of this kind, however, is much less dangerous to the 
attendants, and less costly to repair than the fracture of a cylinder head for instance, 
and for that reason it is not a bad idea to purposely design the piston face weaker 
than the cylinder cover and thus to constitute it a sort of safety-valve. 











146 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

The maximum thickness of the piston barrel (for meaning of symbols see Figs, 
151-154, p. 141) may be made about 

■ 5l= (^ +(+ ' 2 ) ins . (47) 

The decrease of toward the open end depends upon the system of webbing 
used, the position of the rings and of the wrist pin, and upon considerations of 
manufacture. The thickness at the open end may, if the webbing is sufficient, be 
made 

<5o = i<5i to .(48) 

provided there is no danger of drawing the barrel out of round through the wrist-pin 
fastenings (see below). 

To allow for expansion by heat the inner end of the piston barrel, commencing 
at the last piston ring groove, is to be tapered down by about from .2 to .5%. 
The proper amount of taper, that is, the diameter D u can only be determined by 
actual trial, since it differs considerably in different types of engines and with different 
conditions of operation (for instance, size and design of piston, and kind of iron used, 
compression pressure, revolutions per minute, heating value of the charge, temperature 
of cooling water, all have their own effect). Too much play between piston and bore 
leads to leaking, knocking, and incrustation of the piston; too little increases the 
friction, hinders lubrication, and causes scoring of the sliding surfaces. Water-cooled 
trunk pistons are usually only relieved as far as the thickness of the piston face, 
because the clearance between piston and cylinder bore can be easily regulated by adjusting 
the temperatures of the cooling water for piston and cylinder jacket. If outside bear¬ 
ings are used for the piston, the clearance between the latter and the cylinder bore 
is usually made so great that only the piston rings bear against the cylinder wall. (In 
the piston shown in Figs. 164-166, for instance, the clearance is 51.95" — 51.79" = .16".) 

Dished piston faces and edges projecting internally may cause local accumulation 
of heat, which is likely to cause trouble for the first piston rings or the wrist-pin 
bearings. All the rings should if possible be located in that part of the barrel externally 
cooled (by air or by water), and for that reason it is of advantage to make the bridge 
up to the first groove (see Fig. 151, p. 141) equal to 

1.0 to 1.33.(49) 

The thickness of bridge between the grooves, having regard to possible re-turning, 
should always be c^.6. The depth, t, of the grooves should be but little greater, 
up to .02", than the thickness, s, of the rings, especially for the inner rings. The 
clearance between bottom of groove and the rings should be as small as possible, 
because it causes loss of charge or of compression, fills with oil, and the oil, crusting 
down, sticks the rings fast in the grooves. 

To protect the piston rings, and to distribute the wear on the cylinder evenly 
over its entire length, some of the rings are sometimes transferred from the inner hot 
end, to the outer cool end of the piston. This also tends to retain the oil on the 
piston surface in better shape. The snap-rings of vertical engines should be secured 
against turning in their grooves so that the number of joints is evenly distributed 
over the circumference. In horizontal machines it is usual to space all the joints, 





GENERAL ENGINE PARTS 


147 


with proper staggering, over the lower third oi the circumference, or at any rate, 
care should be taken to avoid any chance of the joints registering with the oil grooves 
of the cylinder. (See Fig. 153, p. 141.) 

To obtain uniform distribution of the bearing pressure K the wrist pin should 
be situated in the middle of that part of the piston carrying the load; the relieved 
inner end of the piston should therefore be excluded. Waiving this, the wrist pin is 
sometimes placed nearer the innei end of the piston in order to obtain a shorter 
over-all length of machine. Or again, it is sometimes placed nearer the outer end, to 
obtain a cooler location for the bearing. The latter is compatible with the prihci’ples 
of good design, the former is not. 

In large pistons a proper system of webbing has a marked influence on the 
stiffness and strength. For pistons up to 20" diameter, from 6-8 radial ribs, each 
from to f" thick, together with a couple of strong ribs each side of the wrist-pin 
bosses, is quite sufficient. (See Fig. 160, p. 142.) In larger pistons, at least the radial 
ribs of the lower half of the piston should be extended to the open end, and the 
cross-section of the piston through the wrist pin should be stiffened in addition by a 
couple of circular webs. (See Fig. 155, p. 141.) Thin webs in sufficient number give 
a lighter weight of piston for equal strength than a few thick webs, and by increasing 
the radiating surface, render great aid in the cooling of the piston. 

Water-cooled trunk pistons receive some stiffening through the added walls of the 
water space. But care should be taken to see that the wrist-pin bosses are sufficiently 
stayed with reference to the piston barrel, since no great dependence can be placed 
in the comparatively weak inner jacket wall. The thickness of this wall is usually 
kept down in order to decrease weight and to leave room for the head of the connect¬ 
ing-rod, but it should in any case be sufficient to withstand an internal fluid pressure 
of at least 120—150 lbs. per sq.in. Of course, the pressure of the entering cooling 
water hardly ever exceeds 60-75 lbs., but there is a possibility of the formation of 
water hammer, which would cause momentarily considerably higher pressures. If the 
tightness of the piston face against leakage is to be tested by hydraulic pressure, as 
is specified for cylinder jackets on p. 128, the test pressure to be used determines also 
the proper thickness of metal in the jacket wall. 

The proper lubrication of the piston and its wrist pin is naturally of the utmost 
importance. Oil grooves on the circumference of the piston are of doubtful value if 
the oil is fed merely by drop feed and not under pressure. Such grooves are then 
easily filled up with gummed oil, and become rather harmful than otherwise. If, on 
the other hand, the oil is forced in by a pump at the right moment, the grooves 
are washed clean, and are then of service. In the larger engines only forced feed 
lubrication should be employed, and the oil should be supplied at several points of 
the piston circumference at the same time. This is the only method which insures 
certain, and, at the same time, economical lubrication. The oil grooves should not 
open into the interior of the cylinder. The best place for the exit of the lubricant 
is between the first and second piston rings (counting from the piston face), or even 
one ring beyond this. For the lubrication of the wrist pin in small engines some type 
of drop-feed oiler (supplying the pin by means of a wiper or by a groove along the 
upper piston surface), or a pressure grease cup, is usually satisfactory; but for the 
larger sizes only forced lubrication offers sufficient safety, especially in vertical machines. 

2. The piston rings are usually cast-iron snap-rings of small width, fitted in 
grooves cut around the circumference of the piston. Special spring rings and detach¬ 
able seats for the rings have not proved successful, as nearly all these special con- 


148 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


structions lose their efficiency after a short time by burning fast or otherwise sticking, 
and because they usually require a greater thickness of metal in the piston barrel. 
Rings held out by springs also require a greater depth of grooves, the disadvantage of 
which has been pointed out above. 

In general, the snap-rings are made of uniform thickness. Eccentric rings do not 
possess any appreciable advantage, because the purpose for which they are used, that 
is, to obtain a uniform spring around the circumference, is not realized. On the other 
hand, they cost more to manufacture, and toward the joint the clearance between 
bottom of groove and the ring is seriously increased. 

Constructive Details. Let the external diameter of the snap-ring, when in place, 
be D = 2r ins. (the same as the cylinder bore), let the mean radius of the ring be 

D — s 

r m = ins., and let Kb lbs. per sq.in. be the allowable bending stress. Then 


r D 
s 2s 


4 


K b 

12 p 


+ .5, 


in which p (in pounds per square inch) is the pressure of the ring against the cylinder 
wall. From this the required thickness s of the ring is 


.5D 



in, 


(50) 


The allowable stress Kb in the dangerous section opposite the joint, may be taken 
fairly high, for cast iron from 12000 to 17000 lbs. per sq.in., because the stress is 
constant and acts always in the same direction. Snap-rings, however, are under their 
greatest load when being stripped over the piston to slip them into their grooves. 
This is only possible as long as 


r m D — s^l E 
s ~ 2s = '2.5Kb’ 


(51) 


in which E is the modulus of elasticity. 1 

If it is desired that the stress induced during the slipping on of the rings shall 
be the same as the stress during operation, that is, that a'b—ab, we must have a 
thickness s of ring, assuming £ = 11500000 for cast iron, equal to 


VK b . 

S=r ’”2l45 m ' 


(52) 


From these equations we find the following relation between — and K&, Table 11. 

s 


1 The derivation of these equations is given by K. Reinhardt in the Zeitschrift d. V. D. I., 1901, 
p. 375. Tables 11 and 12 are taken from this very complete exposition of the question of piston 
rings. 











GENERAL ENGINE PARTS 


149 


Table 11 

PROPORTIONS OF SNAP RINGS TO STRIP OVER PISTON 


r 

s 

20 

19 

18 

17 

16 

15 

14 

13 

12 

K b = 

11 600 

12 900 

14 500 

16 350 

18 450 

20 900 

24 300 

28 300 

33 500 lbs. per sq.in. 


The spring pressure, p, per square inch of circumferential r ea, with which the 
ring in place presses against the cylinder wall, varies from 3.5 to 7 lbs. per sq.in. Its 
amount does not appear to depend as much upon the value of the gas pressure to 
be confined as it does upon the rapidity with which the latter rises and falls. 1 For 
any given set of dimensions (see Table 12), 


Kb Kbs 2 

p --— = ■ lbs. per sq.m. 


(53) 


12 - 


6 - 


The part to be cut out of the ring at the joint may be made equal to 


a = 9.5 X ^ 77 — X tt-tv = 9.5rm- Kb 


2 sE 


sE ’ 


(54) 


D-s 

rt = 9.5X-^— XK = d.5r m K .(54a) 

Values of the last term of eq. (54), are given in Table 12, for a series of 

values of — and Kb, assuming the value of £’ = 11500000 for cast iron. 
s 

The width of the ring has no influence on the spring p or the bending stress ob. 
It may vary between 

b — 1 {s and 2s, .(55) 

depending upon the number of rings used. The number of rings may be found 
approximately from 

i>% .(56) 

~ 00 


in which D = cylinder diameter in inches. It is a matter of experience with gas 
engines that narrow rings in sufficient number make a tighter packing with less friction 
than a smaller number of wider rings. 


1 A means often resorted to on testing floors, when the piston blows through, is to conduct the 

pressure of the explosion into the grooves under the rings, to increase the pressure p. Needless to 
say, this means of “last resort’' is wrong from all points of view. 

























150 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Table 12 

VALUES OF V AND k FOR VARIOUS VALUES OF -, —, AND K b> FOR SNAP RINGS 

s s 


r 

s 

r m 

S 

Kb= 

10 000 

11 350 

12 800 

14 200 

15 600 

17 000 

18 450 

19 900 

21 300 

20 

19.5 

V 

K 

2.17 

.0175 

2.49 

.0200 

2.80 

.0225 

3.11 

.0250 

3.42 

.0275 

3.73 

.0300 

4.04 

.0326 

4.36 

.0350 

4.67 

.0376 

19 

18.5 

V 

K 

2.43 

.0166 

2.77 

.0190 

3.13 

.0214 

3.47 

.0238 

3.81 

.0262 

4.17 

.0285 

4.51 

.0308 

4.87 

.0333 

5.22 

.0357 

18 

17.5 

V 

K 

2.71 

.0157 

3.10 

.0180 

3.50 

.0203 

3.88 

.0225 

4.27 

.0248 

4.65 

.0270 

5.03 

.0293 

5.43 

.0317 

5.83 

.0338 

17 

16.5 

V 

K 

3.06 

.0149 

3.48 

.0170 

3.93 

.0191 

4.36 

.0212 

4.80 

.0234 

5.23 

.0255 

5.67 

.0276 

6.11 

.0298 

6.55 
.0319 

16 

15.5 

V 

K 

3.46 

.0140 

3.96 

.0160 

4 43 
.0180 

4.93 

.0200 

5.43 

.0220 

5.91 

.0240 

6.41 

.0260 

6.91 

.0280 

7.40 

.0300 

15 

14.5 

V 

K 

3.96 

.0131 

4.51 

.0150 

5.07 

.0169 

5.65 

.0188 

6.22 

.0206 

6.79 

.0225 

7.33 

.0244 

7.90 

.0262 

8.48 

.0282 

14 

13.5 

V 

K 

4.56 

.0122 

5.21 

.0140 

5.99 

.0157 

6.53 

.0175 

7.16 

.0192 

7.89 

.0210 

8.47 

.0228 

9.11 

.0245 

9.75 

.0263 

13 

12.5 

V 

K 

5.32 

.0114 

6.07 

.0130 

6.83 

.0146 

7.60 

.0163 

8.34 

.0179 

9.10 

.0195 

9.86 

.0212 

10.62 

.0228 

11.35 

.0244 

12 

11.5 

V 

K 

6.27 

.0105 

7.17 

.0120 

8.10 

.0135 

8.97 

.0150 

9.88 

.0165 

10.76 

.0180 

11.65 

.0195 

12.55 

.0210 

13.42 

.0225 

11 

10.5 

V 

K 

7.53 

.0096 

8.61 

.0110 

9.66 

.0124 

10.72 

.0138 

11.80 

.0151 

12.85 

.0165 

13.95 

.0179 

15.05 

.0193 

16.10 

.0207 

10 

9.5 

V 

K 

9.18 

.0087 

10.50 

.0100 

11.80 

.0113 

13.12 

.0125 

14.40 

.0138 

15.60 

.0150 

17.00 

.0163 

18.35 

.0175 

19.70 

.0188 

9 

8.5 

V 

K 

11.50 

.0079 

13.12 

.0090 

14.75 

.0101 

16.40 

.0112 

18.00 

.0124 

19.70 

.0135 

21.30 

.0146 

23.00 

.0157 

24.56 

.0169 


Examples. 1. A piston whose diameter is D = 35.5" is to have snap rings, which, when 
stripped over the piston, shall not have a stress in outer fiber exceeding iv6 = 18450 lbs. per sq.in. 

From Table 11, when iv6 = 18450, the ratio —=16. 

s 

T 

From Table 12, for K b = 18450 and — =16, p, the pressure exerted by the ring, =6.41 lbs. 

s 

per sq.in. Hence the thickness of ring will be 


.5 D 


s = 


Kb 

\l2 p 


+ .5 


.5X35 

48450 


\ 76.9 


.5 


+ .5 


1 . 10 ". 


If the width b is chosen = 1.25 s = 1.375", the number of rings required will be 


35.5 

5X1.375 


= 5 or 6. 


From Table 12 we also find the value of k to determine the part of ring to be cut out at 
the joint. From eq. (54a) we have 


a = 9.5 X — 2 ~ X /c=9.5 X 3 °' 5 0 1 ' 1Q X .026 = 4.25" 


2 
















































GENERAL ENGINE PARTS 


151 


2. Suppose that, for the same piston, the specifications call for a value of p= 5.7 lbs. 
per sq.in. and a value of Kb = 14200 lbs. when the ring is in service. 

T 

Then from Table 12 the value of — will be about 15, that is, the thickness of the ring 
will be 


r 



2X15 


35 5 

The width of ring is 1.25X1.18 = 1.475", and the required number is i =-— -^-—=5. The part 

5 X 1.475 

cut out of the circumference will be, again using k from Table 12, 


QC C _ IIS 

a = 9.5 X ' ' X .0188 = 3.06". 


If 


it is desired to strip these rings in place, the value of — should be, from eq. (51), 


W-. 


11500000 


= 18.0. But the above dimensions show a value of — = 


r m 35.5—1.18 


= 14.5; hence 


'2.5X14200 .. s 2X1.18 

stripping is no longer possible unless there is no objection to Kb exceeding 14200 lbs. per sq.in. 
Another way of solving the difficulty, besides furnishing detachable fastenings for these rings, 
would be to use a cast iron of greater ductility, that is, smaller value of E. 


The spreading of the rings preparatory to stripping them 
into place must be carefully done, otherwise even a properly 
made ring may break. With rings of large diameter, spreading 
is possible only by means of special appliances. Fig. 172 shows 
one of the types proposed by Reinhardt. This is so designed 
that the angle between the two forces tending to spread the 
ring is always about 60°. The greater this angle, the greater 
the danger of fracture, a fact which should be kept in mind 
when spreading by hand. 

To determine the shop dimensions for the ring, express the 
gap distance a, see Fig. 173 b, as a function of the cylinder 

diameter, say a=xD,* or z=-^. Then for the preliminary turning 

make 


the external diameter 


the internal diameter, 


Da — 

Di = 


[« 


1 + .12J 

in., 

[»( 

K! 

| —2s— 

• 12 ] 



Fig. 172. — Spreader 
used to put snap 
ring in place. 


(57) 

(58) 


After cutting out the gap a [a determined from eq. (54) and (54a)], the external 
•diameter of the compressed ring will then be 


D 0 = Da7t r --=(D + .12) in.(59) 


The mean allowance for finishing is therefore .12". This factor disappears in eq. (58) 
if the rings are not to be finished on the inside. Large piston rings require a some¬ 
what greater allowance, while for small rings a somewhat smaller allowance than .12" 
is sufficient. 










152 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Example. 3. For the rings computed in Example 1, the gap a=4.25", hence x = 4.25-^35.5 
.118. Then 

Do=35.5fl+—)+.12 = 36.96"; 


2>i=35.5(l+—) -(2X1.10)-.12=34.52"; 


Z) 0 = ————-^^ = 35.62"= (2) + .12) in., as called for. 

Notes regarding the manufacture of piston rings: The rings are turned and cut off 
from a piece of cast-iron pipe having sufficient wall thickness. Stress should be laid 
on obtaining the proper grade of iron, and it is a good idea to have the pipe cast 
solid at one end. 

The work proceeds about as follows: (see Fig. 173) 

(а) Turn the pipe to the proper diameter, inside and out, leaving a certain . 

allowance (see above) to give the finished ring the required spring when com¬ 
pressed. 

(б) Cut off the rings at once to the required w T idth, so that possibly only slight 
scraping will be necessary to make them fit the grooves closely. 

(c) Cut out of the circumference a certain distance or gap a, see Fig. 1736, so 

that the ring may be compressed to the diameter D 0 . For the ordinary lap joint, the 

piece a may be cut out by milling (Fig. 173a above), or by drill¬ 

ing (Fig. 173a below). Compressing the ring to the diameter D 0 , 
however, gives it a slightly elliptical form, which, in the case of 
small rings, may be taken care of in the allowance made for 

finishing. In large rings, however, this eccentricity must be re¬ 
moved by hammering the inner surface of the ring. This pro¬ 

cedure at the same time increases the spring of the ring. The 
surfaces of the joint in contact should be scraped to a fit. 

( d ) Join the ends of the ring. This may be done as in Fig. 173 c, by passing 
a pin c through the two ends held the proper distance apart, or if the joint is of 
the simple diagonal type, two pieces of thin sheet steel may be riveted over the cut. 

In both cases the method of fastening should be somewhat flexible to prevent any 
straining or warping of the ring. 

( e ) The rings are next turned to the desired outside diameter, and also inside 
diameter, if required. This may be done by turning them one at a time or by holding 
a number of them in a suitable jig. For very important work the rings should not be 
finished at one cut or at one setting; instead, a number of smaller cuts should be 
taken and the hold loosed after every cut to allow of equalization of any possible 
casting stresses. Rings which must be very accurate (for high pressure work) should 
also be finished one at a time. 

(/) After the removal of the joint connection, file out the clearance a', Fig. 

173 c, to an amount varying from .04 to .20", depending upon the diameter of the 
ring. 

The step joint, Fig. 173, is only used in rings which have a diameter exceeding 8" 

and a width of at least . For smaller rings the joint is simply straight diagonal 

or straight across. 




rr^acu 

“1 

<5ooodg! ] 



Li L - 

—i_1 

a! 


p u 

jp 3 

Fig. 

173. 













GENERAL ENGINE PARTS 


153 


The diameter of the wrist pin, sometimes called piston pin, depends upon the 
moment 

Mb=^-~ = pzX.785 Z) 2 X^ in.-lbs.,.(60) 

and 

Mb £ i 3 Kb in.-lbs., 

where d is the diameter of the pin, for see Fig. 154, p. 141. 

For given values of d, l 0 , and P z , the bending stress in the pin therefore is 

10 Mb 10XpzX.785 D 2 Iq 1.964p z D 2 / 0 
<Jb = - ( p = - 4^3 - = ^-lbs. per sq.m., . . . (61) 

and from this, putting Kb= 12000 lbs. per sq.in. for steel, the least diameter of the 
pin must be 

, 3 jl.964p z D 2 Iq 

d = \ - Kb - = \-60or m . (62) 

For our normal diagram, with p z = 356 lbs., 

d = \I^TT in .(62n) 

Eqs. (60) to (62) depend upon the assumption that the pin is supported in its bearings 
with some flexibility and that the load is concentrated at the middle of the pin. In 
fact, the load distributes itself over a length of the pin, which in the most favorable 
case may reach the entire bearing length l. In such cases the bending moment 
reduces to 

Mb in.-lbs.(63) 

Bach, for instance, uses this formula for cross-head pins. In view of the fact 
that, in new machines, the pin hardly ever bears over its entire length, and that any 
deformations of piston or pin bearing will destroy this condition, if it ever exists, 
eqs. (60) to (62) should have the preference. 

The length of the pin depends upon the unit bearing pressure, which, at the 
moment of explosion, should not exceed 1800 lbs. per sq.in., and should be smaller if 
possible. Since 

rr -P z 7 >zX.785 D- 1U 

K ”“=dT = -1- lbs . (64) 

the length of the pin should be 

li * -fX-ragm. (65) 

dK max dii. max 

and with p z = 356 lbs. per sq.in., and A ma x = 1800 lbs. per sq.in., 

.16 D 2 . 

l> — 3 — .. 


d 


. (65n) 



















154 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


In large engines with high explosion pressure, the room available for the inner 
connecting-rod end is in nearly all cases so limited as to necessitate the use of unit-bearing 
pressure greater than the above (up to 21-2200 lbs. per sq.in.). Under such condi¬ 
tions, however, the wrist pin, even if carefully constructed and well lubricated, will 
always be the most delicate part of the entire machine. In every case the pin should 
be made of the best machinery steel and the bearing surfaces hardened. 

In small engines the pin is held in its bearings, which are nearly always slightly 
tapered, by strong set-screws, and is kept from turning by keys. (See Figs. 151-154, 
p. 141.) For large engines, of course, more elaborate methods than this are required, 
some of which are shown in Figs. 155-161. Although they should insure the greatest 
safety, the pin fastenings should be such as to be easily loosened, and when drawn 
up they should not tend to deform the piston. This should be especially noted in 
key fastenings. 

The dimensions of piston rods may be determined from the generally applicable 
long column formula of Euler, which may also be used for connecting-rods, as will 
be explained more in detail later. The material of construction for piston rods should 
be the best grade of ductile open-hearth or crucible steel. 

Formulae (1) to (3) or (3n), p. 191, may be applied without modification to solid 
piston rods. The length L is figured from the center of the cross-head pin to the 
middle of the length of the piston boss. The factor of safety © = 20, used there for 
other reasons, should not be materially reduced for piston rods since the conditions 
of their operation are no more favorable, and any subsequent returning of the rods 
on account of wear must be allowed for. Ample rod diameters are also desirable 
with reference to stuffing-box construction (see p. 138.) 

Usually the piston rods of internal-combustion engines are hollow to allow of the 
passage of cooling water. The moment of inertia of the rod cross-section is then no 
longer 

but I=^d*-M)~&(d*-dS) in.«.(66) 


where d and di stand respectively for the external and internal diameters of the rod. 
From the general long column formula, the allowable load on a column is 


Pk = x 2 JL lbs . 


(67) 


For £7 = 30000000 and ©=20, the allowable total piston pressure then becomes 


Pz = 9.87 


30000000 
20 2 L 2 


(d 4 -d; 4 ) 


740 000(d 4 —di 4 ) 
- f 2 -lbs. 


( 68 ) 


For our assumed normal pressure = 356 lbs. per sq.in., Pz = 280 D 2 , hence 

^ , 26400.—**) . (6gn) 

In vertical engines, eqs. (68) or (68n) may be directly applied. With the factor of 
safety chosen, 0 = 20, the resulting moment of inertia is generally so large that a 
correction with reference to the additional bending moment, due to weight of rod and 








GENERAL ENGINE PARTS 


155 


piston in horizontal machines, is not necessary. To be entirely certain, however, it 
is well to che^k the deflection, /, according to the formula 


L*§Gs+G k 

J 48 El 


in. for machines with tail-rod, 


(69) 


and 



in. for machines without tail-rod, 


. (69a) 


in which G s is the weight of piston rod, and Gk is the weight of piston including its 
charge of cooling water. / must be kept in any case smaller than the clearance 
allowed between piston and cylinder bore. 

The deflection of the piston rod should be confined to its lowest possible amount 
on account of its effect on the stuffing-boxes. If with the maximum deflection 
allowable, eq. (69) demands too great rod diameters, either the piston must be allowed 
to rest on the bottom of the cylinder, or some such scheme as Kollman’s method of 
centering must be adopted. This consists in giving the rod, when under no load, a 
bend upward during the manufacture equal to the deflection /, so that, when in 
service, the weights G s and Gk draw the rod into a straight line. In the shop the 
rod is given this initial bend by the use of a revolving tool head, holding the rod 
stationary. It cannot be done by the ordinary method of turning between centers. 

For information concerning the methods of fastening the piston on the rod, see 
pages 141 and 142. 


V. Crank Shafts 


Material. Cast steel, for the larger engines only the very best grades; mild crucible 
steel with a tensile strength of at least 70000 lbs. per sq.in., and a ductility of at 
least 20%, or open-hearth steel with a tensile strength of at least 64000 lbs. per sq.in. 
and 18-20% elongation. Of late years, in very im¬ 
portant cases, a high percentage nickel steel is some¬ 
times used. This has a tensile strength of from 
•96000-98000 lbs. per sq.in with an elongation of from 
20-22% (in 8"). 

Allowable stress about 12000 lbs. per sq.in.; with 
excellent material and well defined stress conditions 
this may be increased to 14000 lbs. per sq.in. In 
-crank shafts, however, besides the question of strength, 
the stiffness, that is, safety against bending, is the all-important consideration, and 
the dimensions should be taken accordingly. 

1. General Kinematic and Stress Relations. Referring to the notation of Fig. 174, 
the piston position, or portion of the stroke passed over corresponding to any given 
•crank angle a, may be expressed by 


c max 



1 r 2 


c = r(l—cos a)±L(l —cos/?)~r(l — cos a)sin 2 a, 


( 1 ) 


.or, with a ratio of crank to connecting-rod 



x = r (l-cosa)±L[l-v / l-(>lsina) 2 ]~r(l-cosa±Msin 2 a). . . . (la) 







156 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


In these equations, and in what follows, the positive sign will refer to the outstroke, 
and the negative sign to the instroke of the piston. 

In crank mechanisms in which the center line of the cylinder does not pass 
through the center of the crank-shaft, the stroke of the piston is greater than the 
diameter of the crank circle. If the offset between cylinder and center of shaft = a, 
the stroke of the piston will be 

S=x / '(L-\-r) 2 — a 2 — x / (L — r) 2 — a 2 .(2) 

The increase or decrease of the piston velocity may be found from the expression 


c = v -- 11 ( a ~v sin a (1 ± X cos a),.. (3) 

COS u 

^TTCTi T7Z71 

in which v = -gQ— = = uniform linear velocity of the crank pin. 

The acceleration, or retardation, of the piston from- one velocity C\ to another 
velocity C 2 , in the time interval from ti to t 2 , expressed in seconds, may be expressed by 

C2 — C\ 

b = —~t -~( cos -1 cos 2a) = ra> 2 (cos a± X cos 2a).(4) 

The piston has its maximum velocity when a-!-/? = 90°, at which position 


tan /?= il=y-, 
Ju 


and 


c max — 


V 

COS /?' 


Assuming = we shall then have 


(5) 


C m ax — 1 »02v — 1.6 Cntj 


(5a) 


in which c m , the mean piston velocity 


nS 

30' 


(G) 


The force of acceleration, or the inertia pressure, of the reciprocating parts of an 
engine per square inch of piston, is, when the weight of these parts per square inch 

G 

of piston is expressed by Gq = ^-, 


G G 

Pb = ~b = ruP— (cos a± A cos 2a).(7) 

y 9 

Since the angular velocity of the crank we may write eq. (7), 

OU 

P6 = nr 2 ^(^) (cosa±vlcos2a)~-306^^ 2 (?or(cosa±;cos2a). . . . (7a) 
From eq. (7a) we have for the two dead center positions of*the piston 

p s , = .3° 6 (i) 2 ff 0 »'(l ± i) _.306^ 2 G o r(l ±fj .(7ft) 













GENERAL ENGINE PARTS 


157 


The total weight, Gt, is made up mainly of the reciprocating parts (piston, 
cross-head, and piston rod). Only part of the weight of the connecting-rod (from 
.45 to .55 of its total weight) is included in the weight Gt or Go. The accurate 
determination of the inertia effect of the connecting-rod is best made graphically. 1 


Table 13 

WEIGHTS OF RECIPROCATING PARTS PER SQUARE INCH OF PISTON FOR VARIOUS 

TYPES OF ENGINES 


Kind of Engine. 

G[ 

Weight G lbs. per in. of Piston. 

Single-acting Engines: 

Trunk piston *S ^ 1.75 Z) . . . . 

3.5-4.0 

Trunk piston S ^ 2.0 D . . . . . 

4.2-4.9 

'yi/’ifh prfms-hpffd — 15 to 1 75 D . . 

5.6-7.0 

c m n|l Viinrh.cnppH {Hit.nmohilp pn?inps . 

.35-60 

Double-acting 4-cycle Engines: 

fijnalp pvlinrlpr . 

11.0-17.5 


18.0-25.0 

Double-acting 2-cycle Engines: 

Small H P Kortinff) . 

16.0 

T nrap M OOO H P Simnlpxl . 

21.4 




If pi represents the fluid (gas) pressure behind the piston at any time, the net 
pressure on the piston during the time the reciprocating parts are accelerated will be 
pi — pb, and during the time of retardation it will be pi + pb lbs. per sq.in. But since 
the work of acceleration during one stroke of the piston must equal the work of 
retardation, i.e., the sum of these is equal to zero, the power of the engine is 
independent of pb and only depends upon pi. In moderate speed stationary engines 
the inertia pressure, pb, will not usually exceed from 40—60 lbs. per sq.in. of piston, 
but in high-speed motors it may under certain circumstances be much more than 
this. Fig. 175 shows well what abnormal operating conditions may occur with reference 
to the inertia pressures, when the weight of the reciprocating parts is not well adapted 
to the speeds used. The figures refer to an automobile engine with D = 3|", S = 5£", 

G 0 = .64 lbs. per sq.in. of piston, and ^ = 

The total pressure on the piston at any given time is 


. P = piX.785 D 2 . 

This causes a normal pressure upon the guide (trunk piston or cross-head) equal to 
(see Fig. 174), 

N = P tan /?.(8) 

The pressure along the rod is 

k? 

cos p 


Mollier’s Method, described in Zeitschrift d. V. D. I., U well suited to this purpose. 






















158 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


The difference between P and K is so small with the connecting-rod ratios in common 
use, that they may in general be considered equal. 

The turning force (tangential crank-pin pressure) is 


T—P sin («+ft) 
cos/? ’ 


and, finally, the bearing pressure is 


(9) 




P 

cos /? 


cos (a +/?) cos (a +/?) 


( 10 ) 


The curve, Fig. 176, shows the variation in the value of the guide pressure, N, 
for the expansion stroke, as computed on the basis of the normal diagram, Fig. 33 r 

p. 71, and for the ratio A = 

£2>s/Sf.in£4/.4 


/Sj9.eli>S./sp. in. 


Ibs/d 



For the other three strokes N is 
of course much smaller. Since the 
pressure volume diagrams for all 
explosion engines have about the 
same general shape, Fig. 176 shows 
the general variation in N. As 
will be seen, N max occurs between 
16 and 18% of the stroke, and at 
that point has the value 


N„ 


"'to" P z " • • (lift) 


G 0 = .64 lbs. per sq.in. of piston 


^ — 4 5’ ^ — 3.3o" j S = 5.5" 


< 242.3 


39Q5 


This figure is rounded off, taking 
into account any possible late 
ignition. The mean guide pres¬ 
sure during the expansion stroke is about N m ~h P z . 

The variation in the tangential effort or turning force during 

the compression and expansion strokes may be seen from Fig. 178. 

This has again been based upon the normal diagram, Fig. 33, from 
which the pressure diagram, Fig. 177, has been derived. The inertia 

. curves shown in Fig. 177 were obtained by assuming the weight of 
>sV reciprocating parts G 0 = 4.27 lbs. per sq.in. of piston and a speed of 200 
r.p.m. The inertia forces for the suction and expansion strokes, during which 
they are not taken up by gas pressure, cause an alternately positive and negative 
pressure on the crank-pin, which pressure is considerably greater than that due to 
the suction and exhaust resistances. Hence the effect of these resistances may be 
neglected and the tangential effort diagram confined to the second and third 

strokes of the cycle. Figs. 177 and 178 apply to 2-cycle engines as well as to 4-cycle. 

Complete tangential effort diagrams are given under “Computation of Fly-wheels” 
p. 215. 

From Fig. 178 it appears that T mSLX occurs at approximately 40° crank angle, 

m, . , sin («+/?) . , 

Ihe factor cos p ln e( l- (9) for this crank angle (a =40°) is approximately = .75. 

The pressure still remaining on the piston for the same angle is usually in the 

































GENERAL ENGINE PARTS 


159 


neighborhood of Pt~.6 P z . Hence the maximum tangential pressure may be ex¬ 
pressed by 

7 1 max = -75 Pt = .75X.6 P Z = A5 P-„ .(12n) 

Without correction for inertia, the value of Pt and hence of T max , would be somewhat 
greater, and since when an engine is being started, the inertia forces are comparatively 
insignificant, it will be better for the sake of safety to write in general 

P? = .7P ? , . . . (13n) and T max = .5P;. ..... (14 n) 

The above also makes some allowance for possible late ignitions. 

The shaft bearings are always under the maximum pressure at the moment of 



Fig. 176.—Variation in Pressure on Guide. 



explosion. Hence if the two main bearings are placed symmetrically with reference 

p 

to the center line of the cylinder, the maximum pressure on each will be ^ lbs. 



Fig. 178.—Forming Effort Diagram for Compression and Expansion Strokes. 


The magnitude and direction of Z is, however, modified by the weight of shaft 
and of fly-wheel, and by the belt pull, if this method of transmitting the power 
is used. 

Table 14 gives the complex trigonometric functions used in computing 
piston positions, velocities, pressures, etc., for various per cents of the out- and 
instroke. 










































160 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Table 14 

PISTON AND CRANK TRAVEL, WITH THE TRIGONOMETRIC FUNCTIONS INVOLVED 


1 

2 

3 

4 

5 

6 

7 

S 

Piston Position x. 

Crank 
Angle * 
Outstroke 
a 

Rod Angle, 



sin a X 

cos a ± 

1 Outstroke 

4 in % S. 

1" Instroke 

1 in % S. 

sin (<* + /?) 

COS (a + i?) 

(1 + J COS a) 

1 cos 2 a 

2 

98 

14.8 

3.0 

0.306 

0.952 

0.304 

1.141 

4 

96 

21.2 

4.1 

0.425 

0.905 

0.428 

1.080 

6 

94 

26.0 

5.0 

0.515 

0.857 

0.516 

1.022 

8 

92 

30.2 

5.4 

0.583 

0.812 

0.586 

0.955 

10 

90 

33.9 

6.2 

0.630 

0.775 

0.648 

0.905 

12 

88 

37.3 

7.0 

0.698 

0.715 

0.705 

0.848 

15 

85 

41.8 

7.4 

0.757 

0.654 

0.765 

0.767 

20 

80 

48.9 

8.4 

0.842 

0.540 

0.853 

0.631 

25 

75 

55.4 

9.3 

0.904 

0.428 

0.914 

0.498 

30 

70 

61.5 

10.1 

0.949 

0.315 

0.962 

0.369 

35 

65 

67.3 

10.4 

0.977 

0.214 

0.993 

0.244 

40 

60 

73.0 

11.0 

0.994 

0.105 

1.011 

0.126 

45 

55 

78.6 

11.2 

1.000 

0.003 

1.018 

0.012 

50 

-50 

84.2 

11.3 

0.995 

0.096 

1.014 

0.068 

55 

45 

90.0 

11.4 

0.980 

0.196 

1.000 

0.200 

60 

40 

95.7 

11.3 

0.956 

0.292 

, 0.976 

0.295 

65 

35 

101.7 

11.2 

0.920 

0.390 

0.939 

0.388 

70 

30 

108.0 

11.0 

0.875 

0.485 

0.892 

0.471 

75 

25 

114.6 

10.3 

0.839 

0.545 

0.832 

0.548 

80 

20 

121.9 

9.5 

0.751 

0.660 

0.759 

0.616 

85 

15 

130.0 

8.6 

0.663 

0.749 

0.668 

0.677 

88 

12 

135.3 

8.0 

0.597 

0.802 

0.604 

0.709 

90 

10 

139.2 

7.3 

0.552 

0.834 

0.554 

0.727 

92 

8 

143.6 

6.5 

0.500 

0.866 

0.496 

0.746 

94 

6 

148.5 

6.0 

0.431 

0.903 

0.433 

0.755 

96 

4 

154.3 

5.0 

0.353 

0.936 

0.355 

0.776 

98 

2 

161.9 

3.4 

0.253 

0.967 

0.251 

0.789 

100 

0 

180.0 

0 

0 

1.000 

1.000 

o.soo 


* Crank angle for instroke = 180— a. 


2. Friction Losses in Journals. The friction losses considered in the design of 
bearings and pins should be computed on the basis of the average piston or journal 
pressure for one complete cycle (i.e., for four or two strokes, as the case may be). 
On account of the extreme variations in the working pressures and the uncertainty 
regarding the coefficient of friction existing at any given instant, the allowable 
pressures on the sliding surfaces in contact should be taken at a rather low figure. 
The explosion pressure, p z , only determines the maximum unit pressure k max on the 
journal, and this should be placed at such a figure that the oil film can never be 
completely squeezed out of the bearing. This is satisfied in general as long as 
^max. = 1400 to 1600 lbs. per sq.in.; higher values of k up to 1800 lbs. per sq.in. are 
permissible only in extreme cases. 

In the case of 4-cycle engines the average piston pressure of a complete cycle is 
composed of the average pressures on the four strokes, that is, 

P\ T P% 4 * P‘s + Pa u 
p m = - ^ -— lbs. per sq.in, 


( 15 ) 

























GENERAL ENGINE PARTS 


161 


Pi End p 4 for the suction End exhnust strokes Ere due to the inertia forces which 
the crEnk-pin must take up. For the mean pressures of the compression and expan¬ 
sion stroke, i.e., p 2 End p 3 , this effect disEppears, since the positive pressures due to 
the moving parts during the one half of the strokes are balanced by the negative 
pressures during the other half. Turning to our standard diagram, Fig. 33, p. 71, we 
find pi and p 4 each equal to about 18 lbs. per sq.in., p 2 = 32 lbs. per sq.in., and 
p 3 = 128 lbs. per sq.in. (p 3 — p 2 = pi). Hence 

18+32 + 128 + 18 .. 

Pm — -^--= 49 lbs. per sq.m. ~50 lbs. per sq.in. . . . (16n) 

With a given diameter of cylinder equal to D in., the resulting mean piston pressure 
for one complete cycle therefore is 


P»» = 50X.785 Z) 2 = 39.25 D 2 ~40 D 2 lbs.' . . .(17 n) 

It is quite usual even to-day to compute the work lost in friction using the mean 
effective pressure, pi, or even the explosion pressure p z . Either method is incorrect, 
and the computation so made can at best only give some general idea of the magni¬ 
tude of the friction loss for preliminary estimates. This method of computation would 
mean that 4-cycle and 2-cycle engines, single- and double-acting engines, should all 
have the same crank-pin dimensions, which of course is wrong. 

The general equation for the work lost in friction is 

Ar = kv/u=j l X gjfe" ft.-lbs. per sec.(18) 


Equation (18) contains the coefficient of friction, p, which, depending upon the 
load, quality of the surfaces in contact, and the degree of lubrication, may vary from 
.03 to .10. If, on account of its uncertainly, this factor is entirely neglected (it 
cancels from both sides of the equation), we will have the following general formula: 


. Pnn Pn , 

kv= mrm ft - lbs - per sec - 


(19) 


For the crank pin we therefore have 


. p m X. 785 D 2 n p m D 2 n 

kV ~ 230 1 ~ 290 1 


ft.-lbs. per sec, 


(19a) 


From the value for p-n computed above for the standard diagram, we may further 
derive the special equation 


7 50 D 2 n D 2 n n 

fa “-29or=5 W ft " lbs - per sec - 


(20n) 


(a) The Crank Pin. Without reference to considerations of strength , the minimum 
crank-pin length should therefore be from eq. (19a) 


p m D 2 n 
l ~ 290 kv 


in., 


( 21 ) 














162 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 
or, for the standard diagram, p. 71, we have according to eq. (20n), 


1 = 


D 2 n 

5.8kv 


• ( 21 *) 


For the ordinary construction of steel crank pin, the maximum limiting values of kv, 
as based on experience, are about as follows: 


For Bronze Bearings. For Bearings lined with White Metal. 

Aw£1200 ft.-lbs. per sec., A;v£1500 ft.-lbs. per sec. 

Where very careful construction and perfect lubrication is assured, the value of kv 
for white metal can in case of necessity be increased to kv= 1650 ft.-lbs. per sec. It 
is recommended, however, that this value be used in extreme cases only. 

In order to retain the oil film in the bearing during the moment of explosion, 
we must have 

n V 78^ D 2 

km&x = ~-^1400 to 1700 lbs. per sq.in., .(22) 


or 


280 D 2 

^max ^ 


^ 1400 to 1700 lbs. per sq.in, 


(22n) 


For this reason we should be certain that 

Z = ^00d m ‘’ • * • • (23 ^ or l ~5d m .( 23w ) 

In the case of small and medium sized engines, other features of design, especially the 
crank-pin bearing of the connecting-rod, will generally result in dimensions for which 
kv and k max are smaller, i.e., I is larger than the values above given. 

(&) Crank Shaft or 31ain Journals. The bearing pressure in the main journals is 
determined by the piston pressure, the weight of the fly-wheel, and the pull of the 
belt, neglecting the weight of the shaft itself. The weight of the wheel and the belt 
pull generally affect only one of the bearings directly, but the general practice is to 
neglect this fact also. In the case of vertical engines with shaft above the cylinder, 
the bearing away from the wheel is relieved of a part of the explosion load, P z , by 
the weight of the wheel, G, because weight of wheel and force of explosion are 
opposite in direction (most favorable case). In vertical engines of ordinary design, 
with crank shaft below the cylinder, the directions of P z and G arc the same, hence 
these forces are additive, and the bearing next the wheel receives an extra load due 
to G (most unfavorable case). In horizontal engines, the line of action of G is at 
right angles to that of P z , hence the resulting bearing pressures are relatively less 
than in vertical engines with crank shaft below the cylinder. However, the line of 
action of the bearing pressure due to the gas pressure in the cylinder approaches a 
vertical more and more as the expansion proceeds, and this effect may be emphasized 
under certain conditions by the belt pull. For that reason, the following investigation, 
based on the assumption that the lines of action of P z and G coincide, may be 
applied to all types of construction. 







GENERAL ENGINE PARTS 


163 


Referring to Fig. 179, we will then have 

.5 P„ + (£±2l . 5 P m+ G^±5l 

, a 7zain a n 

kV 1J[ X 60x72 = i;- X 230 = 700 ft '' lbs ' per sec ’> • < 24 ) 

from which the required minimum length of main journal may be determined. 

For engines having a third, or outboard bearing, we have the following relation 
for the second main bearing: 

.5 P m + G—^— 

*k v = -—— X 1700 ft.-lbs.per sec.(25) 

In the case of main journals, as in the case of crank pins, the above limiting 
value for kv, which applies to ordinary constructions, may under favorable circum¬ 
stances be exceeded, and to a considerable extent, as some of the examples to follow 

will show. 

The weight of any belt wheel is usually so small as compared with the weight 
of the fly-wheel, that only the latter need be considered. In any exceptional case 
these two gravity forces must be combined. The 
maximum bearing pressure in the outboard bearing 
is usually kept below 400 lbs. per sq.in. 

3. Strength Computations. Side cranks are used 
but little in gas-engine construction, except perhaps 
for some small and cheap engines for general power 
purposes. The objection to them is mainly based on 
the fact that these cranks are supported on one side 
only. 1 

The usual type of crank-shaft (center crank) is 
given a third, or outboard, bearing when the fly¬ 
wheel is exceptionally heavy, or when it is seated 
any considerable distance from the main bearing. 

Above 30 H.P. the use of an outboard bearing is 
more common. In nearly every case the center line of the cylinder bisects the 
distance between the middle planes of the main bearings, i.e., referring to Fig. 179, 
the distance on each side is equal to \a. The usual proportions are 

a = from 1.8 to 2.2 D, 
ai=from .4 to .5a, 

and the center line of the third bearing is usually at a distance ai+a 2 = from 1.75 to 
2.5 a from the center line of the nearest main bearing. 

As far as the crank pins and the crank webs are concerned, it is usually sufficient 
to check them for a maximum load equal to the explosion load P z (crank angle 
a = 0). A simple comparison of the expansion line of Fig. 33 with the tangential 


1 Translator’s Note. This view is apparently not shared by some American builders, who are 
using the Tangye type of engine frame with its overhung crank. 
















164 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


effort line of Fig. 178, will show that P z decreases at a much faster rate than T 

increases, and since the moment arm ~ for P z is always greater than the arm r for T, 

it is plain that the shaft parts mentioned must be under greatest stress at the moment 
of explosion. This applies even to constant-pressure engines (of which the Diesel is a 
type) with cut-off as late as 12-15% of the stroke. Only the diameter of the main 
journal transmitting the power is determined by the forces existing at the moment 
of maximum tangential effort (a~40°). Nevertheless it is well to check the crank 
webs of the larger machines also in this position of the crank, and for machines 
having a heavy fly-wheel this computation becomes necessary. 
In what follows this double check has been used, in order to 
give a better insight into all the conditions that may prevail. 

(a) Maximum Piston Pressure P z in the Dead Center 
Position. In horizontal engines 1 in which part of the weight 
of the fly-wheel is not supported by an outboard bearing (the 
most unfavorable case) the crank pin is acted upon by the 
following moments: 

A bending moment M bl , due to the piston pressure Pz. 

A bending moment M hi due to the fly-wheel weight G 
and a torsion moment Md, also due to G. 

The moment M bl can at once be determined from the 
pressure P z . The other two are most easily found by use of the reaction forces 
at the bearings. These reactions are as follows (for notation see Fig. 180): 


P* 

Due to Pz, R p =~ lbs.(26) 

Due to G, R g i=G—, and R gll = R 01 +G= ib s .(27) 

a a 'm 

Consequently G + R,,i - R yll = 0. 

Mbi=^ in.-lbs.;.(28) 

Mb i = R {t ^ in.-lbs.;.(29) 

Md = R g ir in.-lbs.(30) 


Since Mbi and Mbi are acting at right angles, the combined bending moment 
will be 

M b = VM% + M% in.-lbs. 

The two moments Mb and Md are next to be combined according to Bach’s formula 
(see p. 168, Table 15). The resulting moment is 

Mr = .35 Mb + .65 V M 2 b + <*o M 2 d in.-lbs. 



1 For modifications in the method of crank-shaft computations for vertical engines, see p. 176. 














The factor ao= , OT . . 

l.oAd 

lbs. per sq.in., hence 


GENERAL ENGINE PARTS 


165 


For good steel Kb~ 12000-13000 lbs. per sq.in., and Kd~ 10000 


13000 n 
a ° 1.3 X10000 ’ 

so that ao may hereafter be neglected. 

From the above, the maximum bending stress in the crank pin will be 


Now, for solid crank pins, 


Hence 


Mr 

° b= w 


lbs. per sq.in. 


™_7rd 3 d 3 

W ~l$2 ~ 10 ' 


10 Mr 

° b ~ d 3 


lbs. per sq.in. 


(31) 


. (31a) 


The following example is inserted to make clear the relation existing between Mbi, Mb 2 , 

and Md. 


Example. 1. The crank shaft shown in Figs. 181 and 182 belongs to a Korting engine 
having a cylinder diameter of 10". The number of revolutions is 160. With an explosion 



pressure of p 2 ~350 lbs. per sq.in., the explosion load is P z = 27000 lbs. The weight of fly-wheel 
=<5 = 2640 lbs. Then, from the above equations, 

i? p = .5X27000= 13500 lbs.; 
i2oI = 2640 ^ =145 ° lbS>; 


Hence 


Rgu = R gI +G = 1450X2640 = 4090 lbs.; 


,, 27000X23.6 

Mbi= -— 


= 160000 in.-lbs. 


23 6 

M6 2 = 1450X—^—=17100 in.-lbs. 


Mb = V160000 2 +17100 2 = 161000 in.-lbs. 


















































166 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Md= 1450X9.37 = 13 600 in.-lbs.; 


Mr = .35 X161000 + .65V 7 161 000 2 +13 600 2 = 161 500 in.-lbs. 

The maximum bending stress in the crank pin in the dead center position is therefore 

161500X10 


4.73 3 


- = 15140 lbs. per sq.in. 


Neglecting the two moments M b2 and Md, the bending stress due to Mbi alone would have been, 

160000X10 


4.73 3 


- = 14960 lbs. per sq.in. 


This example shows that it is usually sufficient to check the crank-pin against the 
load P z only, provided the weight G and the arm a are not abnormally great. If, 
however, M bi and M d are neglected, we may derive the following simple expression: 


P z a P z a\0 2.5 P z a 1U 
ab = Wi = ~dH~ = ~d i lbS ' PCT Sq ' m 


(32) 


Assuming o b ^. K b = 14000 lbs. per sq.in. for good open-hearth steel, the diameter of 
the crank-pin will then be 


d> 


3lP z a2.5 
* 14000 



in 


(33) 


Assuming further that a~1.8 D, which is approximately true for the larger engines, 
and putting p z = 350 lbs. per sq.in. for our standard diagram, we will finally have the 
special equation 


31.785 X 350 XD 2 X 1.8 D 
A = y -- s6 _ 


= 1^.09 D 3 = .448 D~.45 D. 


. (33n) 


The limit of safe stress above assumed, i.e., o b S. 14000 lbs. per sq.in., is apparently 
rather high, but in view of the fact that only the best kind of open-hearth steel 
should be used for crank-shafts, this stress may be considered safe, provided no other 
disturbing deformations occur. In the real case, also, the stresses are not quite as 
severe as the above safe method of computation would indicate. For instance, the 
real points of support of the shaft, as far as the explosion load on the pin is con¬ 
cerned, are not located in the middle planes of the bearings, but somewhere near the 
inner edge. For this reason the real free length concerned in the determination of 
the bending moment is not the distance a, but approximately 2c (see Fig. 179). 
Generally, 2c = .65-.75a, hence the real bending moment is from 25-33% smaller 
than the computation with a indicates. It. should further be considered that the load 
P z acts in a normal direction but a very short time, so that in a sense but little 
time for serious deformation of the metal is available. Lastly, the load P z in reality 
is not a concentrated load, as above assumed, but is distributed over the length of 
pin = 1. With these modifications the bending moment due to P z would then decrease 
from 
















GENERAL ENGINE PARTS 


167 


In a similar manner the bending moments derived below for the crank webs and 
the main journals are in excess of the real stress conditions. All the computations 
are thus based upon the assumption of the most unfavorable conditions that can 
exist, hence the results give shafts which are safe not only against imperfections 
in manufacture and other accidents, but which combine great stiffness with sufficient 
strength. 

The Crank Webs or Arms, at the moment of explosion, besides being under a 
purely compressive stress of minor importance, are also subject to bending and 
torsional stresses. The crank web nearest the fly-wheel or belt pulley is naturally 
under the greater stress. In this web we have the following moments (see Figs. 179 
and 180): 



A bending moment, due to P z , 

= M bl =^-e in.-lbs. 

. • (34) 


A bending moment, due to R gh 

= Mb 2 = R g i(r-^j in.-lbs. . . 

. . (35) 

and 

a torsion moment, due to R g i, 

= Md =Rgi (a — e) in.-lbs. . . 

. . (36) 


The stresses due to these moments are: 

In bending, o bl = lbs. per sq.in. . (37) and ob 2 = ^^ per sq.m. • (38) 


In torsion, 


Md 

r “ps lbs - persq - m - 


(39) 


Combining these stresses in the proper manner, we have: 

Bending stress in the two edges under greatest stress due to o bl +o b2 =o b ; 

Stress at the center line of the wide side of the arm due to a bl and r; 

Stress at the center line of the narrow side of the arm due to a b2 and ti = r^-. 

The bending and torsion stresses in the last two cases are combined into a single 
resultant stress, according to Bach’s equation, 

or = .35<T6 + .65\4x6 2 + 4t 2 lbs. per sq.in.(40) 

This equation may be re-written 

—=.35 + .65\^l +4^—V, 

Ob \<n>/ 

and to simplify the computation of a, the following table gives values for ^ when the 
value of — is known. 








168 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Table 15 


$ H 

ii 

0.1 

0.2 

0.3 

0.4 

0.5 

0.6 

0.7 

0.8 

0.9 

1.0 

1.1 

Or 

1.013 

1.050 

1.108 

1.182 

1.269 

1.364 

1.468 

1.576 

1.688 

1.803 

1.921 

Ob 












Ppt = 

3.7 

5.8 

6.4 

8.7 

9.5 

10.4 

10.8 

11.2 

11.5 

11.8 

11.9 

T 

1.2 

1.3 

1.4 

1.5 

1.6 

1.7 

1.8 

1.9 

2.0 

2.5 

3.0 

Ob 












Or _ 

2.040 

2.160 

2.282 

2.405 

2.529 

2.653 

2.778 

2.904 

3.030 

3.664 

4.300 

ob 












Ppt = 

12 

12.2 

12.3 

12.4 

12.4 

12.5 

12.6 

12.6 

12.7 

12.7 

12.8 


The figures in italics in the lines marked Ppt may be used for interpolation. For example, 

Z . Or 

suppose — = .52, then the corresponding value of — interpolated between the values given in the 
Ob Ob 

fifth and sixth columns of the above table may be found as follows: 

Ppt = 2X9.5 = 19, since .52 exceeds .50 by 2 in the last place. Then 



Example. 2. The right crank web of the crank-shaft, shown in Figs. 181 and 182, is under 
stress due to the following moments: 


Mbi = 


27,000 

—W-x 


7.30 = 98600 in.-lbs.; 


Obi — 


6X98600 

3.5UX5.5 


=8600 lbs. per sq.in.; 


Mbi = 1450(9.37—2.37) = 10150 in.-lbs.; 


°bi — 


6X10150 

3A4X5.5 2 


= 570 lbs. per sq.in.; 


Md = 1450(23.6 - 7.30) = 23 600 in.-lbs.; 


9X23600 

2X3.54 2 X5J5 


= 1540 lbs. per sq.in. 


From the above, the stresses will combine as follows: 

In the two edges of the arm under greatest stress, 8600 and 570 lbs. per sq.in.; 

In the wide side of the arm, 8600 lbs. per sq.in. tension and 1540 lbs. per sq.in. shear.; 


3 54 

In the narrow side of the arm, 570 lbs. per sq.in. tension and 1540 X-^—=925 lbs per sq in 
shear. 5 - 5 

The resultant maximum stresses therefore are: 


Edges of arm: 


Ob\ T obi ~8600 + 570 = 9170 lbs. per sq.in.; 


Wide side h : 


Or — .35X8600+.65\/8600 2 + [4X 1540 2 ]=8960 lbs. per sq.in.; 
o T = .35 X570 + ,65\ 570 2 T[4 X925 2 ] = 1450 lbs. per sq.in. 


Narrow side b: 










































GENERAL ENGINE PARTS 


169 


It is clear at once that the stress in the narrow side of the arm is considerably 
below that occurring in the wide side. The thickness of the arm ( b ), in the direction 
parallel to the shaft axis, should be made quite liberal when the fly-wheel is seated 
on the shaft outside of the main bearings and no outboard bearing is used, because 
the explosion load may cause deflections in the crank arm which at once manifest 
themselves by unpleasant lateral vibrations in the fly-wheel rim. According to a well- 
tried rule of thumb, the crank arm thickness in center crank-shafts should be at least 

b = .6 to .7 d. 

The fate of a center crank-shaft too weak in the crank web is well shown by 
Fig. 183. This shaft was part of an engine of 20 H.P., with crank shaft above the 



Fig. 183.—Broken Shaft from 20 H.P. Vertical Engine. 


cylinder, which made it a comparatively large engine of its type. The lateral vibra¬ 
tions of the fly-wheel rim were at first but slight, but during ten years of operation 
they increased in magnitude to such an extent that the engine could no longer 
be used. 

An investigation showed that the crank arm transmitting the power was ruptured 
through about four fifths of its cross-section near the junction with the main bearing, 
and that only the inner one fifth of the cross-sectional area remained sound. A part 
of the area of rupture had already lost its usual granular appearance and appeared 
polished, showing that friction between the surfaces must have occurred during 
operation. As will be seen from the figure, the fracture starts exactly in the joint 
between the journal and the arm, from which we can conclude without doubt that 
the sudden transition from one surface to the other was at least partly to blame 
for the occurrence. It points out the necessity of a liberal fillet at that point, as in 
the case of all other similar constructions under stress. 






170 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


If the weight of the fly-wheel is supported by an outboard bearing, the moments 
M b2 and M d , due to G, disappear. The stresses in the crank arm are then only those 
due to P z , that is, a certain compression stress and the stress due to moment M bl . 
For the latter, eq. (34) may be rewritten for our standard indicator diagram 


from which 


Mbx = 140 D 2 e in.-lbs., 


(34n) 


840 D 2 e 

° b ~~mr- 


(41) 


The dimensions of the main journal transmitting the load are always determined 
by the maximum turning moment. 

( 6 ) Maximum Turning Moment for a =40°. The piston pressure, Pt, acting in 
this position of the crank (see Fig. 184), causes a pressure T at right angles to the 
crank, and a pressure Z in the plane of the crank. From these the stresses acting 
in the various parts of the shaft may be derived, neglecting 
the moment due to G (see above). According to eq. (13n) 
and (14n), p. 159, P< = .7 P z , and T = .5P Z . Hence the radial 
force Z with a =40° and /? = 7.3°, equals 

Z = Pt cos (40+ 7.3 )-.68 Pt lbs., 

or Z = .68 + .7 Pz = ~.48 Pz lbs.(42) 

Of this amount one half = .24 P z , falls to each bearing. 

The exact computation would have to be made, taking into 
account the reactions, Rt, Rz, and R g , resulting from T, Z, and G, 
but since Example 1 showed that the stresses in the crank pin 
are not seriously increased by the fly-wheel weight G, this computation will for the 
sake of clearness not be made here. 

The crank-pin is affected by the following moments, see Figs. 179 and 184. 

A bending moment due to T, M bl = ^ =^~ = .\2b P 2 a .(43) 

A bending moment due to Z, M bi = =. 120 P z a .(44) 

A twisting moment due to Rt, Md = Rtr = .25 P z r .( 45 ) 

Mbi and Mb 2 act at right angles, hence 

M b = VM b 2 + Mb 2 2 = V / (.125 Pza) 2 + (.120 P z a) 2 .(46) 

From the moments as above computed we will then have the following stresses 
v, v 10Mb., 

bending, a b = —lbs. per sq.m.( 47 ) 

~ . 5 M d 1.25 P z r„ 

I orsion, r = -^ 3 - = —^— lbs. per sq.m.(48) 

The resultant stress will be, as before, 



Or = .35<xb + .65\/<7b 2 + 4 t 2 £ K b . 













GENERAL ENGINE PARTS 171 

Example. 3. For the crank shaft of the Korting engine, treated in Example 1, 
Pt = .7X27000 = 19000 lbs., T = . 5X27000 = 13500 lbs., and Z ='.48X27000= 13000 lbs. Then 

13500X23.6 

Mb, =-—-= . 125 X 27 000 X 23.6 = 80 000 in.-lbs. 

13000X23.6 

Mb* =-—-= .120 X 27 000 X 23.6 = 76 500 in.-lbs.; 

from which Mb = \/S0 000 2 + 76 500 2 = 111600 in.-lbs.; 

also Md = .25 X 27 000 X 9.37 = 63 300 in.-lbs. 


Moment resulting from combination of Mb and Md is 

Mr = .35 X111600 + .65 \ TT1600 2 + 63 300 2 = 123 000 in.-lbs. 

The maximum bending stress is therefore 

10X123 000 1irnnn 

Or — - —— - =11600 lbs. per sq.m. 

4. <3 

The computation carried through for the crank-pin in the dead center position, in Example 1, 
gave (76 = 14960 lbs. per sq.in., neglecting G, which value is about one third larger than that 
found above. 


Foi* the crank-pin computation, as above carried out, the resolution of Pt into T 
and Z is superfluous, since we can write directly 


Ah 


Pta 

~T 


.7 PzCL 

4 


.175 P z a 


(46a) 


The same result may be derived by simplifying the radical of eq. (46). But since 
the values of T and Z must be known to compute the crank webs, it is better to 
resolve Pt at the outset. 

The crank arm next to the fly wheel, if called upon, as usual, to transmit the 
entire power, is under compression due to the load hence the stress is 



lbs. per sq.in 


(49) 


This may, however, in most cases be neglected. The arm is further under a bending 
stress due to R z , for which 


M&i = e = .24 P z e in.-lbs., 


(50) 


Mb i 

° bi %hb 2 


1.44 P z e 

hb 2 


lbs. per sq.in, 


and 


(51) 













172 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


There is also a bending stress due to T, for which 

Mb 2 =T(r~Y^ = .5 in.-lbs., 


and 


_M b2 

° b 2 ~W~ 



lbs. per sq.in. 


(52) 

(53> 


Lastly, there is a shearing stress due to R t , for which 

M d =R t e = .25 P z e, .(54) 


and 


T 


M d 1.13 P z e 


lbs. per sq.in. 


(55) 


These various stresses combine, as on p. 168, as follows 

In two edges of the arm, o bl and o bi . 

In the center of the wide side of the arm, a bl and t. 

In the center of the narrow side of the arm a b and n = t^-. 

h 

The bending stresses o b , and the shearing stresses t, should next be combined accord¬ 
ing to eq. (40), or Table 15, p. 168, into one resultant stress which must remain 
smaller than the allowable bending stress K b for the material. The elastic deforma¬ 
tions of the shaft due to the weight G or the reaction R gll tend to relieve the arm 
next to the wheel (see Fig. 180, p. 164), but the determination of the amount of 
stress reduction due to this cause can not be made with any certainty. 

If the power of the engine is taken off equally on both sides of the crank, a 
condition which can not be maintained with any certainty, each crank arm is subject 
to the stress caused by .5 T, so that M b2 and o b2 will be only one half of what they 
were before. 


Example. 4. Considering again the crank shaft treated in the previous examples, the stress 
conditions are the following: 


_ 1.4 4X2700 0X7.3 
° bl 5.5X3.54 2 


=4100 lbs. per sq.in.; 


Obi == 


3X27000(9.37-2.17) 
3.54 X5.5 2 


=5450 lbs. per sq.in.; 


1.13X27000X7.3 

3.54 2 X5.5 


= 3240 lbs. per sq.in. 


Combining these stresses we have: 

In two edges of the arm 4100+5450=9550 lbs. per sq.in. in tension. At the center of the 
wide side h, 4100 lbs. per sq.in. in tension and 3240 lbs. per sq.in. in shear. At the center of 


3 54 

the narrow side 6, 5450 lbs. per sq.in. in tension and 3240 X^rr^ 2080 lbs. per sq.in. in shear. 

5.50 


The maximum stress in this case, as in the dead center position, is found in the edge of the 
arm. By combining o b and r into a resultant stress o r , for each side of the arm, as per eq. (40), 
we find the maximum stress in the wide side, h =6250 lbs. per sq.in., and in the narrow side 
6 = 6320 lbs. per sq.in. 










GENERAL ENGINE PARTS 


173 


The dangerous section in the main journal transmitting the power is close to the 
outside face of the crank arm, that is, close to the junction of the crank arm with 
the shaft. If no outboard bearing is used, this section is under a bending stress due 
to Z and G, while T causes both bending and twisting. The shearing stresses due to Z 
are of minor importance and may be neglected. 

In the case of horizontal machines 1 the reactions R z and R m enclose an angle 
r =90 -40 = 50° (see Fig. 184). The resultant of these forces may therefore be com¬ 
puted from the formula 

Rres. ~ V^z 2 + Run 2 ± 2 COS fRzRyu, 

in place of using the triangle of forces. When 7 -= 40°, cos 40° = .6428, and 2 cos y = 
1.285, hence for this case 

Rres. = 'SRz 2 +Ryu 2 -1.285 R z R oll . 

Where the computation must be very accurate, however, it should be noted that the 
bending moment due to the fly-wheel in this case is 


— G(a 1 +Ci) — RgnCi. 


The reaction resulting from this is therefore 

R, _ C(«. + ej) - = G a 1 +e k _ . 

ei e x 

This, together with the reaction due to Z and T, put the journal under bending 
stress. If, instead of considering the reaction due to Z and T separately, we again 
introduce the connecting-rod pressure Pt, as was done on p. 171, the resultant force 
may be easily found from the equation, 

Rres . = v Rs 2 + R a 2 + 2 RsRg sin 7.3° = VR S 2 + R g 2 + .254 R,R g , .... (57) 

Pt 

in which R s stands for the reaction due to Pi, and is equal to 


Hence the resultant bending moment is 

Mb res. = Rres. .( 57 a) 

The torsional moment due to T is 

M d = Tr =0.5 Pzr .(58) 


These two moments combine to give the maximum stress in the journal. 

In eq. (57) 7.3° is the connecting-rod angle /?, at the instant of maximum turning 
moment, for which a =40°. Sin 7.3° therefore corresponds to the cosine of the angle 
between R s and R g (90 + 7.3°) in the force triangle. 


1 For modifications regarding vertical machines, see p. 176. 













174 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


From the combined moment 


Mr = .35 Mb res. + .65VM b 2 iea .+Md 2 , 


we derive, as before, the maximum bending stress 


from which, for solid shafts 


Mr,. 

<x r = |pr lbs. per sq.m., 


10 Mr 


lbs. per sq.in. 


The above computation is based on the assumption that the journal is called upon 
to resist the entire twisting moment T r . This is really the case, even when the power 
is all transmitted through one journal, only under the most unfavorable circumstances, 
as, for instance, when the engine is just starting up. In normal operation, the moment 
of inertia of the fly-wheel 


7 = SMrf = 


GD 2 

4 g 


GD 2 

128.8’ 


(59) 


tends to relieve the stress in the dangerous section of the journal, so that the twisting 
moment existing at any time is the difference between T r and the moment due to 
the wheel. But since the shaft must meet the highest maximum stresses that can 
occur, the computation should be made with the entire twisting moment T r . 

The middle cross-section of the journal, assuming the shaft perfectly rigid, is only 
under stress due to the twisting moment T r , while the bending stresses due to G 
and Z cancel each other. In any actual case, of course, perfect rigidity does not 
exist, which means that there will also be bending stresses in that cross-section of the 
journal. Owing to the uncertainty existing with reference to the magnitude of the 
moments involved, an exact determination of these bending stresses is not possible. 
The shaft, however, is safe if the journal is given a diameter throughout its entire 
length equal to that computed above for the diameter near the junction of shaft and 
crank web. 


Example. 5. Take again the crank-shaft of the previous examples. Figs. 181 and 182. 
The reactions. 



19000 „ 

= —^—' = 9500 lbs., 


and R g n= 4090 lbs. 


Since G = 2640 lbs., we have from eq. (56), 


R g = 


2640(13.0+5.5) 

5.5 


— 4090 = 4800 lbs. 


Hence from eq. (57) and those following 


#res. = V 9500 2 +4800 2 + .254 X 9500 X 4800=11200 lbs.; 
M b res. = 11200X5.5 = 61600 in.-lbs.; 


M d = 


27000 


X9.37 = 126500 in.-lbs. 


2 










GENERAL ENGINE PARTS 


175 


Combining Mb Te s. and Md, we have 

Mr = .35 X 61600 + .65v / 61600 2 + 126500 J = 113000 in.-lbs 

From this the combined fiber stress then is 


10 Mr 

° r ~~dF 

u n 


10X113000 

4.33 3 


13800 lbs. per sq.in. 


In case an outboard bearing is employed, the effect of the fly-wheel weight G 
disappears as far as the section at the junction between shaft and crank arm is 
concerned, and d u depends solely upon the piston pressure P z or P t . The length of 
shaft ai+a 2 , between the main and outboard bearings is, however, under stress due 
to the weight of the fly-wheel mainly, but also has to take care of whatever stress 
there may be due to the belt pull. As far as this part of the shaft is concerned, 
the influence of the turning effort on the stresses is of secondary importance as 
compared with the bending stresses due to G and the belt pull, even if the equalizing 
action that the rotating masses possess, owing to their inertia, be left out of account. 

The maximum deflection occurring at the center of the wheel seat may be 
determined from 


/-(<? + 


l 2 + ab n \ a 2 b 2 
Sab {tw J 3Ell 


(60) 


in which l=a + b = distance between the middle cross-sections of the second main 
bearing and the outboard bearing [referring to Fig. 179, p. 163, Z = ai + a 2 , a = ai, and 
& = a 2 , G w is the weight in pounds of a length of shaft equal to l, E = the modulus of 
elasticity ( = 30000000 for O-H steel), and / = moment of inertia of the shaft cross- 
section]. The belt pull may be directly determined from the tangential force 
required to transmit the necessary horse-power at the given peripheral speed. That is, 


P = 


550 N n 


60X550 N n 

7znD s 


11000 N n 
nD s 


■ lbs., 


( 61 ) 


in which N max is the actual maximum brake horse-power, v the velocity of belt or 
rope in feet per second, n = r.p.m., and Z) s = the diameter of pulley or sheave in feet. 
In order to possess sufficient friction to transmit the load, the belt must have a 
certain initial tension 

S =——r lbs.,.(62) 

x — \ 

so that the total pull at right angles to the shaft will be 

4 _p£±|lbs.(63) 

The factor x depends upon the kind of belt or rope and the surface of the pulley 
or sheave, also upon the arc of contact, i.e., upon the ratio of transmission. For 
leather belts and nearly equal diameter of driver and driven pulley, we may take the 
load on the shaft due to the total pull equal to 


A =6.5 to 9 P lbs. 














176 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Since the bending -moment due to G is vertical, while that due to A is in the 
majority of cases nearly horizontal, both are to be combined into a single resultant 
moment. The maximum total stress in the shaft length ai+a 2 finally is that due 
to this resultant moment combined with that due to the turning moment T r . 

It should be noted that the maximum stress in the section under discussion should 
be kept low, in order to prevent any noticeable deflections, which are sure to cause 
hot bearings and lateral vibrations of the fly-wheel rim. Further, it should not be 
forgotten that the shaft at this section is usually weakened by key-seats. 

In vertical engines with crank shafts j j the cylinder, P z and G have 

f the same direction] , . , a , , . ,. , , ,, ,... [unfavorably] 

] , which affects friction and strength conditions . , , 1 • 

l opposite directions j l favorably j 

If no outboard bearing is employed, the bending moment, resulting from the com¬ 
bination of P z and G, in the main journal transmitting the power, is 


Mb = Mbi ± Mbz, 


(64) 


in which the plus sign applies to vertical engines with crank shaft below the cylinder, 
and tho minus sign to engines with shaft above the cylinder. The influence of G 
on the crank pin and the crank arms is opposite to that above indicated, assuming 
that sensible deflections do not occur (see Example 7 below). 



At crank position for maximum turning effort (a = 40° and /? = 7.3°), the main 
journal transmitting the power will be under stress due to a force resulting from the 
eombination of rod pressure P t , and fly-wheel weight G, equal to 

R s = VR S 2 + R 0 2 T 2 R s R g cos 7.3° = VRJ* + R g 2 T 1.984 R s R a . .... (65) 

In this equation, when the shaft is above the cylinder, the plus sign applies, while 
for an engine with the shaft below the cylinder (the ordinary type of vertical engine) 
the minus sign is to be used. This is contrary to eq. (64). 

The determination of the twisting moment M d and all other strength computations 
are made the same as for horizontal machines. 


Example. 6. Fig. 185 shows the crank shaft of a 50 H.P. vertical Banki engine. 
Cylinder diameter Z) = 15.75", stroke S = 23.6", w = 135 r.p.m., p z -570 lbs. per sq.in., and G = 
6200 lbs. On account of the unusually high explosion pressure and the rapid drop of the expan- 







































GENERAL ENGINE PARTS 


177 


sion line of the indicator diagram, it will be safer not to determine the reaction forces, 
etc., from the general equations given, but to determine them accurately from the proper 
diagrams. 

As will be seen from Plate IV, Fig. 16, the maximum turning moment occurs much earlier 
in the stroke than is common (a = 21°, /? = 4.1°); the pressure on the piston at this instant 
is ~.8 Pz, from which we can determine the following forces: 


P, = 570 X.785X15.75’ = 110000 lbs.; 
Pt = .8X 110000 = 88000 lbs.; 


T= 88 000 sin (21 + 4 - 1 ) _37 400 i bs . 
cos 4.1 


Z = 88 000 COS(21+4 ‘ 1) = 80 000 lbs. 
cos 4.1 

Mean piston pressure of a complete cycle, also accurately determined 

Pm = 40 lbs. per sq.in., so that P m = 40X.785X15.75’ = 7800 lbs. 
Bearing pressure due to G 


in II, R a ii =6200 


24.4 


24.4 + 18.9 


= 3480 lbs.; 


in III, Rgui = 6200 


18.9 
24.4 + 18.9 


= 2680 lbs. 


(o) Friction Losses: 


Maximum pressure without 
reference to weight of 
shaft or to belt pull, 


Pmax = 8 2 g = 1300 lbs. per sq.in., Crank Pin; 


„ .5X110000 . .. . T 

A max ~ 7 gg x i 34 = 5Qo lbs - P er s< + in -> Maui Journal I; 

(.5 XI10000)+3480 
7.88X13.4 

2680 


Amax ‘ 


= 555 lbs. per sq.in., Main Journal II; 




Mean pressure for complete 
cycle, 


Km = 


Km = 


Krn = 


5.45X13.4 

7800 

8.27X10.23 

5X7800 


7.88X13.4 
(.5X7800)+3480 


= 37 lbs. per sq.in., Outboard Journal III. 

= 92.0 lbs. per sq.in., Crank Pin; 

= 37.0 lbs. per sq.in., Main Journal I; 

= 69.5 lbs. per sq.in., Main Journal II; 


7.88X13.4 
2680 

K m = - -=37 lbs. per sq.in., Outboard Journal III. 

5.45X13.4 

















178 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

Circumferential speed v and work of friction Kmv. 


Crank Pin, 

Main Journal I, 

Main Journal II, 


8 : 27X7rXl35 = 4 88 ft ec K m v=92 X4.88=450 ft.-lbs. per sec. 
12X60 


v=- 


7. 88X^X135 
12X60 


= 4.67 ft. per sec.; K m v = 37 X 4.67 = 173 ft.-lbs. per sec. 


v = 7 - 88x *X135 = 4 6? ft per gec . RmV = 69 , 5 x 4.67 = 326 ft.-lbs. per sec. 


12X60 


Outboard Bearing III, = ft. per sec,; K m v -37 X3.17 = 120 ft.-lbs. per sec. 


(b) Strength Computations . Dead Center Position: 
110000X35.4 


Crank pin: Mb=- 


968000 ,_ onnl . 

= 968000 in.-lbs., ob = 8 2 ~ = 17200 lbs. per sq.m. 


On account of the very high explosion load, it is well to check the pin for shearing stress. 
This is in this case equal to 

.5X110000X4 


T = - 


53.8 X 3 


1360 lbs. per sq.in. 


The combined stress therefore is 

ar = .35 X17 200 + .65V17 200 2 + 4 X1360 2 = 17 350 lbs. per sq.in. 


Since without z, Ob was 17200 lbs. per sq.in., it is evident from the above that even in this 
case the shearing stress in the pin may be neglected. 


Crank arms: 


Compressive stress a = 


.5X110000 

15.3X4.53 


= 800 lbs. per sq.in.; 


Mb = .5 X110000 X10.32 = 568 000 in.-lbs., 


56S 000X6 
4.53 2 X 15.3 


= 10800 lbs. per sq.in. 


o + <76 = 800 + 10800 = 11600 lbs. per sq.in. 

Junction of Main Journals I and II with the Crank Arms: 

J/& = .5X110000X8.08 = 445000 in.-lbs., ffb = —-=9100 lbs. per sq.in. 

! .88 2 


Shearing stress z may be neglected, as above. 

Position of Maximum Turning Moment (a =21°). 

Crank Pin: The simplest way is to' determine the bending stress from Pt (instead of T 
and Z). Then we shall have 


Mb — 


88000X35.4 


= 780000 in.-lbs., ob- 


780000X10 
8^27 3 


= 13800 lbs. per sq.in. 















GENERAL ENGINE PARTS 


179 


From the turning effort T we have 

Md — —~— XI 1.8 = 221000 in.-lbs., r= —= 1940 lbs. per sq.in. 


Combining ob and t, we have 


Or = 15 740 lbs. per sq.in., 


which is considerably smaller than the stress at the inner dead center position. 
Crank arm transmitting the load: 


From R z \ Mbi = X 10.32=414000 in.-lbs., 

& 


414000 X 6 __„ 

OT,, “l5.3X4.53’" 7900lbs ' perSqm - 


From T: = 37400 


=295000 in.-lbs., =1200 lbs. per sq.in. 


From Rt: Md ^-^ X 10.32 = 192000 in.-lbs., 


X — 


4.53 X18.1 2 
192000 

| X4.53 2 X 15.3 


= 2750 lbs. per sq.in. 


T 3 18700 

From —: Shear T= 2 X 453X153 = 405 lbs ’ per Sq ‘ in * 


From the above the combined stresses are: 

Center of wide side of arm: 

<76i = 7900 lbs. per sq.in., x r = 2750 + 405 = 3155 lbs. per sq.in., 


from which 

Center of narrow side of arm: 


Or — 9300 lbs. per sq.in 


<76 2 = 1200 lbs. per sq.in., Tr = -^ 


2 

+ 405 2 = 910 lbs. per sq.in., 


from which 

In the edge of the arm: 


Or = 1830 lbs.per sq.in. 


<76 = 7900 + 1200 = 9100 lbs. per sq.in., r = 405 lbs. per sq.in., 

from which <j t = 0130 lbs. per sq.in. 

Here again the stress is less than at the dead center position of the crank. 

Main Journal transmitting the Load: 


88 000 

Mb = -r— X 8.08 = 355 000 in.-lbs., 


M d = 37400XI 1.8 = 440000 in.-lbs., 
Combining these gives 


355000X10 ., 

ob = -——-- <260 lbs. per sq.in.; 


x— 


7.88 s 
440000X5 


7.88 3 

<j r = 10100 lbs. per sq.in., 


= 4500 lbs. per sq.in. 


which is greater than the stress found for the same section in the dead center position. This 
example shows, as was stated at the outset, that the crank pin and the crank arms receive their 














180 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

maximum stress through the load P z , while the maximum stress in the main journal transmitting 
the power is due to the moment 7V. 

In vertical as well as in horizontal machines, the stresses occurring in the shaft 
are favorably or unfavorably affected by the elastic deformations of the shaft under 
load, depending upon whether the slope of the elastic curve is directed against or 
with the lines of action of the external forces. In general, the change is to make the 
real stresses less than those computed above on the assumptions set forth. This is espe¬ 
cially true for those parts of the shaft where the maximum stresses may be expected to 
occur, that is, in the crank pin and the crank arm and main journal transmitting the 
load. The method used in the above computations therefore combines simplicity with 
great safety. The exact determination of the various stresses involved in the crank¬ 
shafts, taking into account the influence of deformation in the material, is quite 
complicated, and is therefore made only in very important cases as a check. The 
main advantage of such an investigation is that it gives more exact information as 
to the real pressures on the main bearings than it is possible to obtain with the 
simple but more usual method. 

The approximate shape assumed by the elastic curve of a crank-shaft at the 
moment of explosion is shown, much exaggerated, in Fig. 187. 



Example. 7. Figs. 186 and 187 show the crank-shaft of a 100 H.P. vertical Giildner engine. 
Maximum piston pressure is approximately 125000 lbs., weight of fly-wheel is about 12300 lbs., 
the belt pulley is about 220 lbs. From these weights we may derive the maximum stresses in the 
dangerous sections of the shaft as follows: 


Table 16 



Dead Center Position. 

Crank Position a=35°. 

Crank 

Pin. 

Crank 
Arm Under 
Load. 

Middle 

Bearing. 

Crank 

Pin. 

Crank 
Arm Under 
Load. 

Middle 

Bearing. 

| Seat of 
Fly-wheel. 

Principal stresses, lbs. per sq.in., 
according to 

(а) Simplified method. 

(б) Strict method. 

16 300 

14 400 

10 850 

8 800 

8100 

4850 

10 500 

9 250 

6950 

6100 

10 250 

8 400 

5550 
















































GENERAL ENGINE PARTS 


181 


In the above table the values in line (a) were computed by use of the simplified method 
above developed, while the values of line (b) are based on principles laid down in the well- 
known thesis 1 of Ensslin. 

The latter also furnishes the following values for the maximum deflection at the moment 

of maximum turning effort (<* = 35°): Slope ft = greatest deflection in the shaft 

next to the seat of the fly-wheel is less than .012 in. 

The friction losses in this shaft are as follows: 


Table 17 



Journal. 

Crank Pin. 

I 

II 

III 

Maximum pressure, Kmax, lbs. per sq.in. 

Mean pressure, K m , lbs. per sq.in. 

Circumferential speed, v, ft. per sec. 

Work of friction, K m v, ft.-lbs. per sec. 

427 

47 

6.28 

296 

423 

98 

6.32 

620 

48 

4.95 

238 

1340 

146.5 

6.57 

960 


The design of multiple-throw crank-shafts may be based on the methods of computa¬ 
tion already developed. Since the maximum turning moment Tr, for each crank always 
occurs within the first one fifth of the expansion stroke, no two !T max can occur at 
the same time for any of the multicylinder combinations now used. For that reason 
the diameter of the shaft may in this case also be computed on the basis of one 
T m&y r, as i n the previous instances. As for the rest of the dimensions, the greater 
free length of shaft (due to greater distances between bearings) cause greater bending 
stresses (due to P z in the dead center position), which call for a corresponding increase 
in the dimensions. In the case of engines having four or more cylinders, the phases 
of the cycles in the various cylinders should be so arranged that, in any pair of 
cylinders having cranks 180° apart, expansion does not occur in one cylinder while 
compression takes place in the other. If this point is neglected it may happen 
that the turning moment due to the explosion in one cylinder is accompanied 
by the resisting moment of the compression in the other, which puts the intervening 
crank arm under stresses quite possibly excessive. The phases in the various cylinders 
of a four-cylinder machine, for instance, should be arranged as indicated in Fig. 188. 
The crank arrangement shown there is also the most favorable as regards balancing. 
The inclined position of the middle crank arm must be regarded as a make-shift only 
to which recourse must be had when the cylinders are so close together that there is 
no room for a bearing between them. In general this construction should be avoided, 
because the inclined arms are subject to the action of additional forces which seriously 
increase the stresses occurring, besides complicating the method of computation. The 
following example furnishes evidence regarding these points: 

Example. Figs. 188 and 189 show the dimensions of the crank-shaft of a 4-cylinder 
4-cycle marine engine whose cylinder diameter =9.85"; stroke = 12.55"; and r.p.m. = 350. Taking 
the explosion pressure at 356 lbs. per sq.in., the load on each crank will be P z = 356 X .785 X 9.85 2 


1 Max Ensslin, Mehrmals gelagerte Kurbelwellen mit einfacher und doppelter Kropfung, Stuttgart, 

1902. 

















182 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

~27000 lbs Pt = 7X27000-18000 lbs., the maximum tangential force r = .5X27000=13500 
lbs, and the radial pressure Z = .47X27 000 = 12 700 lbs. The weight of the light fly-wheel may 
be neglected. Assume that the second cylinder from the left is just exploding. Since the shaft 
is symmetrical on botli sides of the middle line only one half needs to be checked. 



Figs. 188 and 189.—Dimensions in Inches 

Crank Pin. In dead center position: Reactions due to Pz 

27000X16.1 


In bearing I, Ri- 


37.7 


= 11500 lbs. 


. „ 27000X21.6 

In bearing II, Rn =--= 15o00 lbs. 


37.7 


Bending Moment on the Crank Pin: 


250000 


Mb =15500X16.1 = 11500X21.6 = 250000 in.-lbs., o b = | x5 ~ 8 = 15100 lbs. per sq.in. 

* 

Crank Position, a =40°. The connecting-rod force Pt causes a reaction in bearing II equal to 

18000X21.6 


Rn — 


37.8 


= 10300 lbs. 


In this position, therefore, the crank pin is under the following stresses: 

i, 166000 u 

Bending due to R u : Mb = 10300X16.1 = 166000 in.-lbs. <7b = i ^r^ = 9800 lbs. per sq.m. 

Torsion due to Rt: M d = 7750X6.3=487000 in.-lbs., r= "2x55^ =144 ° bs ' ^ Sq ' m ' 

Combining ob and r, the resulting maximum stress is a r = 19 100 lbs. per sq.in. 

Crank Arm Transmitting Load. Dead Center Position. In this position the arm is under 
stress due to Ri or Rn only. Hence 

Mb = 15500XI 1.90 = 186000 in.-lbs., ob= 6 7 X3 ' (j r =18b00 lbs ‘ per sqjn * 














































































GENERAL ENGINE PARTS 183 

Crank Position a = 40 ; The reactions due to T and Z furnish the following stresses: 

Bending due to R z : Mbi = 7280 X11.90 =87000 in.-lbs., obi = — 000X6 = 8700 lbs. per so in 

6.7 X 3.0 2 1 1- 

Bending due to T: Mb 2 = l350oU.3-~\ =48000 in.-lbs., ( 7 b z = ^ Q ° X6 =2170 lbs. per sq.in. 

Torsion due to Rt: Af d=7750 X 11.90 =93000 in.-lbs., ? = -■■ ■ = 7000 lbs. per sq.in. 

9 Xb./ X o.U 

Hence using Table 15, p. 168, we shall have the following combined stresses: 

In the wide side of the arm 

From <76! = 8700 lbs. per sq.in., and - = 7000 lbs. per sq.in., a r = 13800 lbs. per sq.m. 
In the narrow side of the arm 

3 0 

From (762 = 2170 lbs. per sq.in., and T! = 7000X=3130 lbs. per sq.in., c/ r = 4900 lbs. per sq.in. 
In the edges of the arm 

C76 = (76i + (762 = 8700 + 2170 = 108"0 lbs. per sq.in. 

Again the maximum stress is considerably less in this position than in the dead center 

position. 

The same method of computation would be applicable to the intermediate crank arm, if it 
had its normal position at right angles to the axis of 
the shaft. For the inclined arm, on the other hand, the 
method is as follows: 

Intermediate Arm. Dead Center Position. The first 
step is to resolve the reactions at the bearings due to P z 
into components parallel and at right angles to the 
center line of the inclined arm. This is most easily 
done graphically by means of the force polygon. Fig. 

190 shows the method and needs no further explanation. 

It is sufficient also in this case to confine the computa¬ 
tions to one of the arms. The dangerous sections to be 
considered are those at x-x and at y-y. 

The moments acting on the former section are: 

M'bi = 14700X20.8 = 306000 in.-lbs. Fig. 190.—Dimensions in Inches. 

Opposed to this moment, however, is that due to the component of P z at the crank pin, 

which is 

M"bi = 25400 X 3.55 = 90000 in.-lbs. 

Hence the net moment is 

Mbi = 306000 -90000 = 216 000 in.-lbs. 

A second pair of moments acting on this section are 





and 


M'bi =9200X11.8 = 109000 in.-lbs., 
il/"6 2 = 5300X11 = 58000 in.-lbs. 

















184 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


These again are opposed to each other, and the net moment is 

Mbz = 109 000-58 000 = 51000 in.-lbs. 

Now Mbi and Mb 2 also act in opposite directions, so that the net effective moment on x-x 
finally is 

Mb = 216 000 - 51000 = 165 000 in.-lbs. 

The section modulus of the cross-section x-x, referred to the wide side of the arm as the 
base, is in this case iX6.7X3.82 2 =16.1 in 4 , so that the maximum stress in this section is 

165000 „ • 

( 76 = =10200 lbs. per sq.m. 

Taking next the section y-y, we find the following: 

Moment M'bi =14700X20.8 = 306000 in.-lbs 
Moment M"bi = 25400 X 3.55 = 90000 in.-lbs. 

Net moment Mbi — 306000 — 90000 = 216000 in.-lbs. 

Further 

Moment M'b 2 =9200X3.94 = 38000 in.-lbs. 

Moment M"b 2 = 5300X3.15 = 16500 in.-lbs. 

Net moment Mb 2 = 19 500 in.-lbs. 

Hence net effective moment is 


d/6 = 216 000 -19500 = 198500 in.-lbs. 

Since the section modulus remains the same, the maximum stress in the section y-y is 

196500 

(76=—— = 12200 lbs. per sq.m., 

which is considerably greater than the stress in x-x. 

Crank Position a = 40°. For this the reactions due to T and Z are determined, and resolved 
into components parallel with and at right angles to the crank arm under discussion as above. 
The stresses resulting from these various forces are less than those shown above for the dead 
center position, as was proven by computation. For this reason the figures are not repeated 
here. 

Main Journal. The computations for the main journal are made as previously outlined and 
offer nothing new. 

Designs of Crank-shafts. 

























































GENERAL ENGINE PARTS 


185 


Figs. 191 and 192. Crank-shaft of a single-acting blast-furnace gas engine. D = 52", S = 
55.2", n = 90 r.p.m. Weight of fly-wheel about 168000 lbs., of the shaft, including counter¬ 
weights, 48400 lbs. Explosion pressure normally 25 atm. = 356 lbs. per sq.in. 

The checking through of this shaft gives the following results: Maximum beating pressure 
on the crank pin, A max = 1100 lbs. per sq.in.; in each of the bearings I and II, A mi , x = 460 lbs. 
per sq.in. Maximum pressure in the bearings due to the weight of fly-wheel and shaft, in 
bearing I, K = 17.8 lbs. per sq.in.; in bearing II, A= 135 lbs. per sq.in., and in bearing III, 
K = 106 lbs. per sq.in. The mean piston pressure, P OT = 106000 lbs., is distributed as follows: o~ 
, . . „ 106000 

the crank pm, A = — 2 x27 6 = 152 bs ‘ per sq-in -’ m each of the hearings I and II, K= 6^.2 lbs. 

per sq.in. From these figures the average bearing pressures for one complete cycle are: for the 
crank pin A = 152 lbs. per sq.in.; for bearing I, A = 17.8 + 63.2 = 81.0 lbs. per sq.in.; for bearing II, 
A = 135+ 63.2 = 198.2 lbs. per sq.in., and for the outboard bearing III, A = 106 lbs. per sq.in. Per¬ 
ipheral speed of the crank pin is u = 9.84 ft. per sec., of each of the three bearings, v = 7.8 ft. per 
sec. Finally, the friction work for the various bearings is: for the crank pin Kv = 152X9.84 = 1485 
ft.-lbs. per sec.; for bearing I, Kv = 81 X7.8 = 632 ft.-lbs. per sec.; for bearing II, Kv = 198.2X7.8 = 1550 
ft.-lbs. per sec., and for bearing III, At’= 106X7.8 = 826 ft.-lbs. per sec. 

The following table shows the stresses computed on the basis of the method developed 
on p. 164, etc. 


Dead Center Position: 

at = 13100 lbs. per sq.in., crank pin; 
ob= 9100 lbs. per sq.in., second crank arm; 

Cr = 6050 lbs. per sq.in., second main journal. 
For the position of maximum turning effort (a =40°): 

Or = 9800 It>s. per sq.in., crank pin; 

Or— 6830 lbs. per sq.in., second crank arm; 


o r = 7450 lbs. per sq.in., second main journal. 


The maximum deflection of that part of the shaft carrying the wheel is in this case less than 
.0006". 


-M 06 - 




[Figs 193 and 194.—Dimensions in Inches. 


Figs 193 and 194. Crank-shaft of a double-acting two-cylinder 4-cycle engine.. D = 25.2", 
8 = 30 7" and n = 150 r.p.m. The assu med unit explosion pressure of 356 lbs. per sq.m, shows a 
maximum bearing pressure on the era ik pin of 1160 lbs. per sq.in., and on the main journal of 
3 S4 lbs. per sq.in. Since the weight of fly-wheel is not given, a complete examination of this 

shaft cannot be made. 
















































186 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

Figs. 195 and 196. Triple-throw crank-shaft of a six-cylinder single-acting marine engine 
(Loutzki). Normal rating 300 H.P. Assuming p z = 356 lbs. per sq.in., we find the maximum 
bearing pressure for the crank pin = 1240 lbs. per sq.in.; for the inner main bearing it is 850 lbs. 
per sq.in. The velocity of rubbing is, in the one case, 10.8 ft. per sec.; in the other, 10.3 ft. per sec. 


///. - -Q05 -*j 



Figs. 195 and 196.—Dimensions in Inches. 


The design of the crank mechanism must finally take into account the question of 
balancing. This is especially important in the case of high-speed engines, or those for 
which sufficiently rigid foundations cannot be had. The forces to be balanced are (a) 
those due to the centrifugal effect of the crank arms and that part of the connecting- 
rod considered as a rotating mass, and (6) those due to the inertia pressure (see 
p. 156) of the reciprocating parts. The latter force acts in the direction of the 
center line of the cylinder. There is no difficulty encountered in exactly balancing the 
effect of the rotating weights (a) by other rotating weights having the same static 
moment. The reciprocating parts (6) of weight Gt pounds, applied at the crank radius r 
(see p. 156), require for perfect balancing a counterweight G s , applied at an arm n, 
and diametrically opposed to Gt, such that Gtr=G 8 r x . 

The counterweight G s (assuming the connecting rod of infinite length) causes a 
centrifugal force equal to 

C = '° 31 GsV - = (.00034 G s r x n 2 ) lbs.(66) 

n 

Of this force the component acting in the direction of the cylinder axis for any given 
crank angle a is 

Ci =C cos a = (.00034 G s r x n 2 cos a) lbs.(67) 

This force balances the inertia effect due to Gt. The second component of C, that at 
right angles to the axis of the cylinder, is 

C 2 = C sin a = (.00034 G s r\n 2 sin a) lbs.(68) 

This force, C 2 , in the case of horizontal machines, is taken up directly by the founda¬ 
tion, but in vertical machines its action across the frame may set up serious vibrations. For 
this reason it is usual in stationary vertical machines to balance only the rotating parts when 
any balancing is done at all. The axial inertia forces in vertical machines really compel the 
use of balance weights only when the effect of these forces approaches the weight of the 
engine. The weight of stationary vertical machines is usually from 100 to 150 lbs. per sq.in. of 
piston and that of high-speed auto-engines from 3 to 12 lbs. Hence the inertia forces are 
always less than the weight of the stationary vertical engines, but for the automobile engines 
the reverse is usually the case. Now since the body of automobiles is much more strongly 
affected by vertical than by horizontal forces as regards vibration, it pays, in the case of 
vertical auto-engines, to completely balance both the rotating and the reciprocating masses, 
irrespective of the action of the lateral force C 2 . 







































GENERAL ENGINE PARTS 187 

Some designers of automobile engines use the following empirical formula, taken, as far 
as the author is aware, from French automobile practice* 

G *= p i + ~ J ^ K lbs.(69) 

In this formula P i +P^ = P = the weight of the connecting-rod, K = the weight of the 
piston. Pj is the weight of the crank pin end of the rod determined when the rod is 
supported on a knife-edge at a distance equal to one third of the rod length from the 
crank-pin end. P § = P-P 4 is the remainder of the rod weight. It is said that 
vertical single-cylinder engines are nearly perfectly balanced if the counterweights are 
determined according to this formula. 


Example. A six horse-power automobile engine has a cylinder diameter equal to 4.33", 
stroke=4.73"; n = 1200 r.p.m. The weight of the connecting rod = P 4 +Pg = 3.88 lbs.; of the 
piston is 5.08 lbs. Supporting the rod, on narrow side, upon a knife-edge one third of the 
distance from the crank-pin center, and resting the crank-pin end on scales, the weight shown 
is P$ = 2.16 lbs. Hence P§ =3.88—2.16 = 1.72 lbs. counterweight, if applied at an arm equal to 
the crank radius, should therefore weigh 


ft . 2J6+ t5+!y»-3» ib. 


The correctness of this computation was checked very closely by supporting the engine shaft on 
knife-edges. 


A theoretically perfect balancing of the inertia forces is rendered impossible by the 
finite length of the rod. This causes an unequal distribution of these forces over the 
two halves of the stroke, while the action of the transverse component due to the 
counterweight is always symmetrical. Taking the best case, i.e., when Gtr = G s r\, the 
amount of vibration in a machine balanced to this extent compared to that found in a 


/ 

machine entirely unbalanced may be expressed by the relation j-: 


1-+ r 


With f =-i 
L 5 


as an average case, there would then still remain a force equal to ^C a = lC a . If, on 

the other hand, the static moment of the counterweight is only mGtr, in which of 
course m< 1, the above relation is rendered less favorable according to the ratio 

(— ■ 


If possible the counterweights should be placed in the planes of the crank arms. 
Counterweights in the fly-wheel cause additional twisting forces, because they are out 
of the plane of the rotating and reciprocating masses, and their use should be avoided, 
especially in case no outboard bearing is used. The room available between the main 
bearings is usually insufficient to admit of counterweights large enough to completely 
balance the reciprocating parts. It is usual therefore in vertical engines to balance only 
the crank arms and that part of the rod considered as rotating. In horizontal engines, 
also, the reciprocating parts are balanced as far as possible; in most cases, however, 
only about one half the weight necessary for this can be placed as indicated. 1 2 


1 Compare Radinger, Schnell-laufende Dampfmaschinen. 

2 See also p. 156. 






188 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


VI. Connecting=rods 

Material. For the body of the rod usually soft steel; for the larger engines the 
use of cast steel is on the increase, while for the cheaper grade of small machines 
malleable iron is sometimes employed. Crank-pin bearings, when the pin diameter 
exceeds say 5", are usually of cast iron or cast steel lined with white metal. For the 
crank-pin bearings of smaller engines and for the wrist-pin bearing the use of bronze 
is nearly universal. 

In computations regarding the strength of the rod body considered as a long 

column, the usual factor of safety is © = 20. 

Crank-shafts for internal-combustion engines in most cases admit of tw 7 o part 
crank-pin bearings only. Among these the so-called marine type of bearing seems to be 
the best adapted; the strap or stirrup type, on account of its lack of sufficient rigidity, 
should be used for the smaller sizes of engines only. For smaller rods the crank-pin 
bearing is also often made as a separate piece, the material being bronze, cast steel or 
cast iron (Figs. 197-200 and 205-206). One-piece construction of crank-pin bearing 
and rod body (see Figs 201-204 and 207-224), however, is to be preferred for the 
larger engines, because this arrangement assures greater stiffness and relieves the bolts 
of the inertia forces acting at right angles to the path of the rod. The wrist or 

piston -pin bearing is usually a solid head with bronze steps. White metal linings 
have usually not proven themselves well adapted for service in this end of the rod. 
Their application, however, is not entirely without promise of success, since of late 
years the piston-pin bearings of some of the large gas engines are lined with white 
metal and have so far given no trouble. In many of the smaller engines the wrist-pin 
bearing is designed without any arrangement for taking up wear, the bearing being 
simply a solid head with a bronze or hardened steel bushing. Rods with forked wrist- 
pin ends, Figs. 217-224, are used in connection with cross-heads only. In the larger 
sizes of trunk piston, the wear in the wrist-pin bearing should be taken up by drawing up 

that half of the bearing not under load, that is, the inner half. Owing to the lack of 

room, however, and the general inaccessibility of this bearing, this procedure is not 
easy with the usual methods of construction. Special constructions with this aim in 
view are shown in Figs. 225-229. In these the inner step may be drawn up by adjusting 
the bolt in front of the pin. Connecting rods with separate crank-pin heads may be so 
made that the rod length may be varied in order to change the compression, Figs. 
197-200 or 205-206. This is of special advantage in experimental machines. 

Design of Connecting Rods. 

Figs. 197-200.—Connecting Rod 
for Giildner 2-cycle Engine, 
Fig. 35, p. 86. 

(Body of rod cylindrical, as 
as it passes through a stuffing 
box. The brass crank-pin bear¬ 
ing is merely clamped on the 
rod body in order to be able to 
adjust the length of the rod or 
the size of the clearance space.) 


































































GENERAL ENGINE PARTS 


189 



Figs. 291-204.—Connecting Rod for a 6 H.P. Banki 

Engine. 

(With reference to mass-production and cost of manu¬ 
facture, the rod body is made practically rectangular. 
The wide sides of the rod are either planed or milled, 
the narrow sides are milled, no lathe work being done.) 


Figs. 205 and 206.—Ciank-pin End 
of Connecting Rod for a 40-50 
H.P. Hornsby-Akroyd Engine. 
(Plates a are used to adjust the 
length of the rod, and con-equently 
also the compression space, to suit 
the kind of oil used.) 



Figs. 297-209.—Connecting Rod for a 20 H P. Banki Engine (using water injection). Built by Ganz & 

Co., Budapest. 

(Standard construction for rods for all of the larger engines ) 



Figs. 210-216.—Connecting Rod for 100 H.P. Guldner Engine (D = 18.75", S = 27.5"). 

(The rod body is cast steel, the crank-pin bearing cast iron lined with white metal, the wrist-pin 
bearing is bronze.) 































































































































































































































o' jf 


n 


A/ 1 , 

\V^-1— 1 

i 

i 

«-i 

. 

A 


Vi 




F 3z': 

“T 

- 

1 


o 







7 --§h 


Ht 

•T-f 




- r 
i 




"7 


Figs. 219-224.—Connecting Rod for a Large Gas Engine with Cross Head. 

Pz 

(Concerning the forces —, P s and P n , see p 193.) 



Connecting Rod. 

(The cover plates a-a surround the inner, 
unloaded half of the bearing, which may be drawn 
up by wedge b. Plates a are pressed firmly against 
the sides of the head by means of the wedge 
screws and are thus held in position.) 



Rod. 


(The bearing is in two parts, block a and wedge b are 
located in that part of the bearing not under load.) 

190 

















































































































































































































































































GENERAL ENGINE PARTS 


191 


Constructive Details. The strength computations mainly concern themselves with 
designing the rod body as a long column safe against failure under the load P z . Care 
should be taken, however, to see that the crushing stress in the smallest rod section 
does not exceed the safe limit. The bending stress resulting from the friction in the 
bearings and the weight of the rod may be neglected (see Example, p. 193). In large 
engines with high rotative speed, the inertia forces in the plane of the rod cause 
additional bending stresses which should be computed. 

The general long-column formula of Euler (for a round-ended column) is 

= lbs. (breaking load).( 1 ) 


Putting E = 30 000 000 lbs. per sq.in., the smallest moment of inertia of the rod body 
for the explosion load P z must be 

7 -4 

1 300000000 m ' ’. (la) 

where G = factor of safety. 

Placing 6 at 20, we have from eq. (1) the allowable load on the rod 


73 7i 2 X30000000X/ ., mnnnn / ,, 

P z =- 20^2 -15000 000 -^2 lbs. 


For a circular cross-section 


7r# 

64 * 


(2) 


Hence, for a round rod, the required mean diameter of the body is 

~l±PzI? • 

150000007r ^738000 m ‘ ' * 



From this, with p 2 =356 lbs. per sq.in., and P z = 280 D 2 , for the standard diagram, we 
finally have 


•4 


T) 2 L 2 
2640 = 


4 


DL . 
52 11 


. . (3 n) 


Toward the wrist pin the diameter decreases to from .7 to .75 d m ; toward the crank 
pin it increases, depending upon the requirements of the crank-pin bearing; if necessary, 
the width required for the bearing at this end may be obtained by flattening the rod 
laterally (see Figs. 207-209). 

For a rectangular rod cf mean height h and width b, the moment of inertia 
(still considering the rod as a pin-ended column, that is, failure occurring in the plane 
of the rod) is 

I = T \h 3 b. 

Hence for this case 


„ 7T 2 X 30 000 000 X h 3 b 1250000 h 3 b „ 

Pz = - ^ o--lbs. 


20X12L 2 


L 2 


... (4) 
















192 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


The value of h is usually = 1.7 to 2 b. Substituting in (4), ^ = 1.856, and solving for b, 


or, for the standard diagram, 




4 PzL 2 


7950000 




D 2 L 2 


IDL . 


28400 ' 169 


( 5 ) 

(5 n) 


In the usual constructions b is kept constant throughout the length of the rod. The 
factor of safety used, © = 20, is a little smaller than is commonly employed in 
stationary steam-engine practice, but is quite safe. Even a further reduction of © is 
permissible, since the maximum load P z acts for a very short time only, which tends 
to prevent serious deformation. For the connecting-rods of small high-speed engines © 
is sometimes decreased to 5, to keep the inertia forces of the rod down to the lowest 
practicable limit. 

The maximum bending moment, in the plane of the motion of the rod, caused by 
the inertia forces is, according to Bach, 


Mb = .000002n 2 rA yL 2 in.-lbs., 


( 6 ) 


and the resulting stress therefore is 

L 2 

tf6 = .000002n 2 7vl7'^r lbs. per sq.in.(7) 

In these equations 

n = revolutions per minute; 
r = radius of crank in inches; 

A = area of mean section of rod in square inches; 
y = weight of rod material in pounds per cubic inch; and 
L = length of rod in inches. 


Putting W = \bh 2 we shall have as a special equation applying to rectangular rods, 


or since A=bh, 


Now the requirement is that 


^ = .000012n 2 rA r ^ 


*6 = . 000012 n 2 ry^- .( 8 ) 


a b + -j- = °b + O Kb, 


hence, for the case of the rectangular rod, 


P L 2 

Kb^.^ + .0000l2n 2 ry~ lbs. per sq.in. 


(9) 











GENERAL ENGINE PARTS 


193 


The moment of inertia of the I section commonly employed for cast-steel rod bodies is 

BH 3 — bh 3 

1 = - yi -• In rods of this section the material is utilized to better advantage than 

is the case in round or rectangular rods, and these rods are consequently lighter. For 
cast steel, E may be taken at an average equal to 31000000. 


Example. Engine dimensions are D = 37.5", 5=51.3", n = 125 r.p.m. Here 


P z = 280 X37.5 2 = 396000 lbs., and L is taken at 5r = 5X ~^ = 128". 


For a round rod: 


-4 


, 1396 000 X128 2 _ 

73800 q ~ 9 -' ~ 9 4 • 


For a rectangular rod, mean thickness: 


396 000 X128 2 


4/39600 
' 7 95 


950 000 


= 5.35" ~5f' 


The mean height of the rod then is 


h = 1.85X5.37 = 9.93" ~10". 


In the case of the round rod the area of the mean section is 74.57 sq.in., for the rectan¬ 
gular rod it is 53.7 sq.in., which shows that for the same service the latter type of rod is 
somewhat lighter than the former. The inertia forces cause a bending stress in the rod of 

ob = .000012 X125 2 X25.6 X .29 X—= 2300 lbs. per sq.in. 


This stress is considerable, but the stress does not exceed the allowable, for since 


396000 

5.37X10 


= 7400 lbs. per sq.in., 


ffmax = a + Ob = 7400 + 2300 = 9700 lbs. per sq.in. < Kb. 


The forked wrist-pin ends of connecting rods, Figs. 217-224, receive their greatest 
stresses in the section x-x or y-y. In the first of these we have the crushing stress z 

.5 P z 
° = l)h’ 

and the bending moment 

il/b = .5 Pze. 

In the section y-y we have the shearing force P s , the twisting force P n , and the 
bending moment Mb = .5P z z. These separate stresses thus appearing are to be 
combined as previously shown. 











194 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


In single-acting engines the caps and bolts of the connecting-rod bearing are only 
under the load due to inertia of the reciprocating parts and the friction of piston and 
cross-head. This load in stationary engines does not usually exceed 10 lbs per sq.in. of 
piston. As far as the cap is concerned, therefore, the size required to hold the bearings 
is of more importance than the stresses it has to resist. It is also necessary to make 
the diameter of the bolts greater than this comparatively small load calls for, in order 
to compensate for any stress due to the strong drawing up required and to allow for 
the inertia effect of the rod in its plane of motion. Weak crank-pin cap bolts are a 
very serious danger during operation, and this particular machine part calls for ample 
dimensions as well as for care in the selection of material and in manufacture. (Soft 
O-H steel, no sharp grooves or corners, smallest bolt cross-section outside of thread, 
safety against working loose, etc.) In double-acting engines of course both the 
connecting-rod bolts and the crank-pin cap must be designed for the maximum piston 
load P z . 


VII. Valves 

Material. For valve cages, or valve seats, hard close-grained cast iron; for valve 
disks mostly medium carbon steel or nickel steel; for small automatic valves sometimes 
also soft steel; large valves, especially those water-cooled, are made with cast-iron disks 
and stems of medium carbon steel. 

The allowable stress in valve disks should be very low on account of high tempera¬ 
ture, to guard against warping, and to furnish material for re-grinding. Numerical 
data is given in the computations to follow. 

The valves of the working cylinders of gas engines of to-day are exclusively of 
the poppet type. The inlet valves may be automatic, but the exhaust valves must be 
mechanically operated. Slide valves, on account of the usual high explosion pressure 
and temperatures, can only be used in places where their dimensions are small and 
their stroke short, as ignition or starting valves, for instance. 
For any other purpose they are obsolete. For all important 
valves in regular and continuous operation, the vertical form 
of valve should be used. Horizontal valves are much more 
liable to leakage, sticking, and other interruptions of regular 
operation and are not reliable especially when automatic. If 
the use of the horizontal form is unavoidable, make the disk 
as light, the guide as long, and the lift of the valve as 
short as possible. And generally, in the case of horizontal 
valves, flat seats are better than the usual conical seat of poppet 
valves. 

Automatic inlet valves are applicable to small and comparatively slow-speed engines 
only. At the higher speeds the accompanying noise becomes very annoying, the wear 
is increased very seriously, and the volumetric efficiency is low. For these reasons 
mechanically operated inlet valves have of late years been used even in small auto¬ 
engines. In the larger machines the use of the automatic inlet valve is prohibited 
alike by decreased economy and capacity and by general unreliability in operation. 
The weak point of all automatic valves is rapid wear of the disk and liability to frac- 



Figs. 230 and 231. 
Fastening of Spring Plate. 













GENERAL ENGINE PARTS 


195 


ture. With this type of valve, screw fastenings are almost useless, and consequently 
the spring plates and other parts, in order to stay in place for any length of time, 
must either be pinned or keyed in (see Figs. 230 and 231), or they must be in one 
piece with the stem. The breaking away of the disk from the stem may, in the case 
of vertical engines whose valves are placed in the head directly over the piston, 
become positively dangerous, and safety appliances for the catching of loose disks 
should therefore be provided in such cases. 

It is advantageous to place the exhaust valve disk in a water-cooied valve housing, 
or in the water-cooled cyUnder head, so that the valve can be cooled by the jacket 
direct. When separate uncooled valve seats, as bushings or cages, are employed the 
proper cooling of the valve disk is much more difficult. In single-acting horizontal 
machines, the inlet valve should, if possible, be placed directly over the exhaust valve, 
because the incoming new mixture will help to keep the exhaust valve cool, and the 
inlet valve works easier if opening downward. Since the exhaust valve disk at the 
moment of opening the valve is loaded with from 30-60 lbs. per sq.in., the reaction 
upon the valve gear when large valves are used may be quite serious. To avoid this 
difficulty, double-seated valves have been devised which are partially balanced either 
by the pressure of the exhaust gases themselves or by compressed air furnished for 
the purpose. To gain the same end, two valves instead of one are sometimes employed. 
Another scheme is to use one valve and an auxiliary exhaust port. When two valves 
are used, one should open a little ahead of the second for the purpose of relieving it. 
For the same reason, in case the auxiliary port is used, the piston should start to 
uncover this before the exhaust valve begins to open. By means of this preliminary 
exhaust the exhaust gases of highest pressure and temperature are discharged in a very 
satisfactory way, leaving nothing but low pressure, and hence comparatively cool gas, 
for the exhaust valve to handle. 

The exhaust gases, owing to their high temperature, have a very serious effect 
upon the valve disk, proper lubrication is very difficult, and hence above a certain 
size, say 100 H.P., direct water cooling of the disk must be resorted to. (For 
example see p. 199). It has been attempted to combine the inlet and exhaust valves, 
and thus not only to keep the combined valve comparatively cool by the incoming 
fresh charge, but also to simplify the mechanical details of the machine; but none of 
these forms, of which Figs. 250 and 251 give some examples, have been able to 
maintain themselves. If they are to be at all reliable, their construction is not 
cheaper but more costly than that of two simple valves, especially when the inlet is 
mechanically operated. The sucking back of exhaust gases from the valve housing can 
hardly be prevented, and the double valve is too vulnerable against soot, rust, etc. 
Such three-way valves, invented and exploited time and again therefore appear to the 
experienced designer to offer little promise of utility. 

The main aim of gas and mixing valves is to produce as far as possible a uniform 
mixture of gas and air, and this determines the general features of their design. The 
means employed to this end are: Increasing the opportunity for diffusion by dividing 
the stream of gas into many fine streams or broad thin layers, proper guiding of the 
various gas currents, mechanical agitation, etc. In the case of automatic mixing valves, 
care should be taken to see that for all valve positions the ratio of the areas of gas 
and air inlet-ports shall be the same, otherwise the composition of the mixture depends 
upon the piston speed and the frictional resistance of the valves. Positively operated 
mixing valves in general give a better guarantee of constant mixture. For gases not 


196 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


entirely free from impurities (suction gas), automatic mixing valves are not satisfactory 
in any case. 

The starting valves, found in all the larger engines, operate only for a few turns 
of the engine at a time, and hence its construction is not of very great importance. 
However, to avoid any trouble due to this inactivity, the valve should be so placed 
that it will keep fairly cool and that no chance be given for sticking it by deposits 
of burned oil, etc. This valve may be actuated entirely by hand or temporarily and 
positively by a part of the valve gear. It is self-evident that the main valves should 
all open inward, and the same is true of the starting valves, even if this is constantly 
loaded by compressed air under a pressure greater than the maximum working pressure 
p z in the cylinder. A common form of starting valve shows a design something like 
that of an ordinary globe valve, that is, the disk is screwed down on its seat. This 
form is not recommended, because the explosion pressure comes upon the rather weak 
valve stem and easily causes the valve to leak. Besides this, the contact surfaces are 
easily corroded and encrusted. 


Design of Valves. 

For forms of ignition valves, see p. 293, for starting 
valve, p. 251. 



Fig. 232. —Inlet and Mixing 
Valve for Small Engines. 



Fig. 233. —Inlet and Mixing Valve combined with 
Governing Valve, Brombacher Type. 



Fig. 234.— Mixing Valve, Koerting Bros. 
(Automatic disk valve with dash-pot at top). 



































































GENERAL ENGINE PARTS 


197 



Figs. 235 and 23G.—Mixing Valve, 
Guldner, used mainly on the 
large vertical engines. 

(Disk a controls the gas ports b by 
means of the ports a', while ports a" 
regulate the air supply. Bell c helps to 
make the mixture uniform. The lift of 
the valve is adjusted through d. 
Throttle valve e controls the air supply.) 


Fig. 237. —Inlet and Mixing Valve for a large Simplex Engine, 
(a air, b gas, c inlet valve, all operated by d d'.) 
Concerning the governor details e-i, see p. 243. 



Fig. 243.—Exhaust Valve with Water- 
cooled Stem-Guide, used by Maschinen- 
bau-Ges. Nurnberg for Medium Sized 
Engine. 

























































































































































































































































































































































198 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 



Fig. [244.— Water-cooled and balanced Exhaust Valve for a 
large Crossley Engine. 

(a main valve, b auxiliary valve. This valve is operated by c and opens 
just ahead of the main valve. This allows the pressure in the cylinder to 
act on the underside of o' and thus relieves the main valve.) 



showing Spring 
H.P. Niirnberg 


&e _>) Cast /ron 


■ 

—t ata. Ac, at 


Hardened , 

Steet Steet 


Tension Piag 
> tor Ttain 
_]) ‘Spring 


am Spring 


;t; J4 Turns. Extension 
~P \4SS,m to a tengfhof 
Spring erguat to j (2.S in. 

Figs. 245-249. — Water-cooled Exhaust Valve, 
Suspension and Guide Construction, 600 
Blast Furnace Gas Engine. 



haust Valve. 


(The stem of the valve a carries a con¬ 
centric sleeve b. The latter, for maximum 
lift, closes off the upper inlet ports, while for 
the smaller lift it opens the exhaust ports.) 



Fig. 251. — Combined Inlet and Ex¬ 
haust Valve. 

(Valve a and concentric sleeve valve b are 
operated independent of each other. During 
suction the valve b has, in its lowest position, 
closed off the exhaust ports, while during 
exhaust, b in its highest position, resting 
against c, has closed off the inlet ports.) 

A valve of similar construction is shown in 
Part III in connection with the Loutzki 
automobile engine. 


























































































































































































































































GENERAL ENGINE PARTS 


199 



Fig. 252.—Exhaust Valve, Pawlikowsky type. 

(Water-supply pipe a is stationary and surrounded by the 
valve spindle. The latter is bored out large enough so that the 
hot water can find its way out through the annular space be¬ 
tween the inside wall of spindle and the water-supply pipes. 




Fig. 254.— Exhaust Valve, Pawlikowsky Type, for 
100 H.P. Korting Engine. 


Fig. 253. —Exhaust Valve, Giildner Type with Pawli¬ 
kowsky Method of Water Supply. 


































































































































































200 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Constructive Details. Let C s be the volume in cubic feet of gas to pass through a 
given valve per second, v the velocity of the gas in feet per second, h the lift of the 
valve , and d the diameter of passage, both in feet. Then the free cross-section /', 
required through the valve, in square feet, will be 

f'=7tdh = ^ sq.ft.,.(1) 


if the lift in every case is For the maximum effective lift h—~, we ma y write 

C s = .785d 2 v cu.ft.,.(2) 

and from this the required minimum diameter of the valve passage will be 


= feet..(3) 

- \785v v v 


But in operation the introduction of a charge through a valve occurs intermittently 
with varying velocity, hence we are compelled, in determining the free cross-section 
required, to base our computations on the true interval of and the quantity of gas 
necessary for a single stroke. Assume that a volume of gas equal to the entire stroke 
volume = Vh cu.ft. (without reference to the volumetric efficiency Tj e ) is to pass through 
the valve and let the diameter of piston be D", the area of piston be F sq.in., S the 
stroke in feet, and c feet per second be the mean piston speed at n revolutions per 
minute. Further, let the clear diameter of the valve seat be d", the lift of the valve 
be h", and f sq.in. be the free cross-section neglecting the area of the stem. Then if 
v ft. per sec. is to be the mean velocity of the gases passing the valve, we must have 


1447* 

v 


Fc 

=-— sq.m, 
v 


(4) 


And, again, providing that h is always we may write eq. (4), since c=^, 


, „ ~D 2 Sn 

/-*»--Y-jgj, 


„ D 2 Sn . 
dh = sq.m. 


120y 


(4a) 


The mean charging and discharging velocity v should not exceed 75 ft. per sec. (in 
large engines this may of necessity go as high as 100 ft. per sec.) and substituting this 
figure in eq. (4a) we finally have 


dh = 


D 2 Sn 

9000 


sq.in, 


(5) 


According to page 156, a ratio of connecting-rod to crank equal to 5 makes 
Cmax = 1.6c, so that if v is taken at a mean of 75 ft. per sec., the maximum velocity 
of. passage through the valve will be 120 ft. per sec. To keep within this maximum 
velocity limit, any given piston position requires a valve lift of 


> D 2 Sn 
~9000d 


in. 


(6) 















GENERAL ENGINE PARTS 


201 


In this equation the factor ^=sin a (1 ± /. cos a) expresses the variation of piston 

velocity when the connecting-rod ratio £ = X. For the usual ratio, L = 5r, Table 18 
gives the numerical values of ip. 


Table 18 

VARIATION OF PISTON VELOCITY (ip) FOR >1 = 1:5 


Part of outstroke. 

.02 

.04 

.06 

.08 

.10 

.15 

.20 

.25 

.30 

.35 

.40 

.45 

.50 

Part of instroke. 

.98 

.96 

.94 

.92 

.90 

.85 

.80 

.75 

.70 

.65 

.60 

.55 

.50 

^» = sin a(l±A cos a). 

.304 

.428 

.516 

.586 

.648 

.765 

.853 

.914 

.962 

.993 

1.011 

1.018 

1.014 

Part of outstroke. 

.55 

.60 

.65 

.70 

.75 

.80 

.85 

.90 

.92 

.94 

.96 

.98 

1.00 

Part of instroke. 

.45 

.40 

.35 

.30 

.25 

.20 

. 15 

.10 

.08 

.06 

.04 

.02 

0 

^ = sin a(!T;A cos a). 

1.00 

.976 

.939 

.892 

.832 

.759 

.668 

.554 

.496 

.433 

.355 

.251 

0 


Plotting the values of ip from the above table as ordinates and the corresponding 
piston positions as abscissae, we shall obtain a sine 
curve similar to the piston velocity curve, Fig. 174. p. 

155. From this curve we can at once determine 
the necessary valve lift and the outline of the actua¬ 
ting cam. If proper co-ordinates are chosen, the h 
curve, Fig. 255, will lie outside of the ip curve through¬ 
out, but this can only be done when the valve starts 
to open slightly before the beginning of the stroke, and 
closes slightly after the end. 

The maximum lift — can be adequately utilized only in the smaller moderate speed 

machines. In large engines questions of valve design, and in high-speed engines the 
proper instant of closing of the inlet valve make it necessary to keep the maximum 

lift much below 4- Automobile engines usuallv show h=~ to for automatic inlet 
valves. 

If there should be more than one valve, each is to be designed as above indicated 
for the amount of gas it is called upon to handle. In hit-and-miss engines care should 
be taken to see that the valve held open during the misses has sufficient free opening 
in this position. 

The dimensions of the valve cages, guides, or retainers, if they are used, depend 
upon the general conditions of the particular design. To make gas-tight joints between 
these parts and the cylinder heads or cylinder walls into which they are fitted, metallic 
packing rings or surfaces ground conical at an angle of 60° should be employed. If 
the cages are fairly large the force with which they must be pressed home to secure a 
tight joint is considerable, and if the conical joint is employed there is therefore 



















































202 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


danger of splitting the housing. In such cases, therefore, it is better to use flat 

surfaces with metallic packing rings. 

As far as the dimensions of the valve disk are concerned, 
besides d or d 0 (Fig. 256) only 8 can be determined by com¬ 
putation. For mechanically-operated valves with disks of medium 
carbon steel we may make 



6400 

or putting p z = 356 lbs., 


._J p.(-5do) a _ ; .(7) 

\ funn ’ . 




^ in .(7 n) 


18 


To derive this equation the disk has been considered a flat plate supported at the 
edge, and in view of the high temperature and the fact that an allowance should be 
made for regrinding, the allowable stress has been taken as Ab = 5700 lbs. per sq.in. 
For the purpose of decreasing the weight of an automatic valve, Kb may be increased 
up to 11000 lbs. per sq.in. In the larger disks, say d^. 4", 8 may be decreased to 
from i to §d toward the edge. 

The width b of the valve seat may be made approximately equal to 


b = .5(do-d)=.0ld + 1 \ in.. • (8) 

For automatic valves this should be made a little greater. The angle of the seat is in 
most cases about 45°. 

For the diameter of the valve stem, the following formula gives satisfactory figures: 

s = + to -g-d + yV in.(9) 

The stem of the exhaust valve may be made a little greater in diameter than this in 
order to decrease the wear and to facilitate conduction of heat to the cooling water. The 
length of the guide depends primarily upon the general design of the valve cage or 
housing, and does not depend directly upon the size of the valve. It is evident that 
short valve actuating levers and considerable frictional resistances require longer guides 
for the valve stem than if the reverse were the case. To prevent possible jamming or 
tipping up of the valve disk, the guide should be extended if possible to the point 
where the actuating lever takes hold. But even this is not sufficient for the exhaust 
valves of large engines, for which a separate guide to take up the lateral pressure of 
the valve lever should be provided (see Fig. 252). Other points which require careful 
design in connection with large exhaust valves are the cooling and lubricating arrange¬ 
ments for the valve stem guides. 

The diameter d 2 of the valve housing may be found from the requirement that 


.785(di 2 — do 2 ) > .785d 2 .(10) 

This is satisfied as long as 

di>1.6d.(11) 
























GENERAL ENGINE PARTS 


203 


It is generally the case, however, that the form of the combustion chamber furnishes 

more than ample area in the annular space between disk and housing. 

The valve springs are generally cylindrical, spiral steel springs. For these some 
general information will be given at the end of Part H, page 312. The requisite 

tension in these springs depends mainly upon the vacuum occurring in the cylinder. 

For mechanically-operated valves, the effect of this vacuum may be taken equal to a 
load of from 6 to 7.5 lbs. per sq.in. of cross-section of valve; for automatic valves 
these figures change to from 11.5 to 14.0 lbs. per sq.in. The tension of the springs 
should be such as to be able to overcome the suction action of the cylinder under all 
conditions of operation, to prevent accidental sucking back of air or exhaust gases. If 
the valve springs should also be employed to help fetch part of the valve gearing back 
into certain positions, the inertia of such parts of course calls for a corresponding 
increase in the power of the spring (see p. 205). It might be said that this method of 
actuating part of the valve gear through part of its movement is no longer good 
practice in the larger machines. 

In the larger sizes of engines not only the load on the springs, but also their 
deflection in operation, are considerable, and for that reason care should be taken that 
the spring is not made too short. In the position of maximum deflection, that is, ■ with 
the valve wide open, there should be from .05 to .10" play between the individual 
coils. The number of coils and the length of the spring when free should be computed 
with this allowance. The following example shows the method: 


Example. A given inlet-valve, opening downward, has a diameter in the clear of 4.73" 
and a lift of 1.18". The weight of the disk including spring plate, nut, and stem, is 14.5 lbs., 
which corresponds to a load of .85 lbs. per sq.in of passage. Maximum 
vacuum in the cylinder is 7.1 lbs., hence with closed valve the minimum 
tension in the spring should be Py = - 785X4.73 J X (7.1 + .85) = 139 lbs. 1 
When the valve is wade open, the tension P 0 will be from one third 
to one half greater than P g . Now turning to Table 39, p. 312 the 
spring w r hich seems to fit both the load and the disk best is one wdth 
a diameter of 2f" (measured as usual from center of -wire) and with 
thickness of wire (<J) = .218". For this spring the table gives the maxi¬ 
mum load P m ax = 185 lbs., and a maximum deflection /, for every turn, 
equal to 1.08". If a spring of f = 10 turns be used the maximum avail¬ 
able deflection will therefore be 10.8". Now, from a diagram, Fig. 257, 
drawn wdth P max as abscissie and / ma x as ordinates, it is at once seen 
that in order to obtain a spring tension of 139 lbs., the spring must 
be compressed an amount equal to 2=8". This must be the initial 
deflection of the spring when placed on the valve stem. The diagram 
also shows that by compressing the spring equal to the required lift 
h = 1.18", the maximum tension in the spring wdll be Po = 162 lbs. (wdiich satisfies the requirement 
that Po^Pmax)- The theoretical length of the free spring should be at least 

l'=id+z+h = ( 10X. 218)+8.0+1.18 = 11.36 ins. 

With this length, however, and the valve open equal to the lift, the coils of the spring would 
be in contact, which is not permissible. To make some allowance for increase of initial com- 


Fig. 257. 



1 Where the computations must be extremely accurate, the effective disk diameter, i.e., up to 
the middle of the seat, should be used. In Fig. 256 this is d + b instead of d. 













204 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


pression of spring, should this be found necessary, the clearance between the coils should be 

d 

made ample, say e=—. The required length of the spring then will be 

O 


l=id+e + h+i- = (10X.218)+ 8.0+ 1.18 -H10X— 1=12.08, say 12 ins. 1 


/ .218\ 
(10X-): 


The accurate determination of the tension of valve springs is especially important 
in the case of the automatic inlet valves of high-speed engines, since this tension 
determines the suction resistance and the volumetric efficiency r) £ (p. 31) of the 
cylinder. Our present knowledge of the mode of valve action does not enable us to 
derive a rigid mathematical determination of the most favorable load on the valve 
disk. In view of this, the most rational method is to regard the valve disk as a body 

G 

of weight G, or mass which must be accelerated during the time t, and 

through the distance h( = valve lift). The general equation for this case is 


1 


Ph=—mv 2 =— 


1 G 


2 32.2 


or 


p . G 

Pt = mv= 325 v - 


( 12 ) 


Putting ^ = 32 ^ = \ 32 ~ 2 i we find the mean spring tension required in automatic inlet 
valves, for the time and lift chosen, is 


and inversely, 


2 mh_ 2 Gh Gh 

t 2 32.2f 2 ~16f 2 lbSv 


(13) 


M6 P f 


0 K ^lGh 
= .25sec. 


(14) 


Equation (13) gives the required average minimum tension in the spring. Friction has 
here been left out of account. In the case of valves opening downward it may be 
assumed that any guide friction is balanced by the weight of the disk. 


VIII. Valve Gearing 

Material. For cam or lay shafts, machinery steel, cams and rollers, hardened steel; for 
the valve levers soft steel or cast steel; for driving gears generally, steel upon cast iron or 
upon bronze in the case of screw gears, and cast iron upon cast iron for spur (bevel) 
gears. For low intensity of pressure and low velocity, satisfactory service is also given 
by cast-iron worm gears, if constant lubrication is provided. 

The allowable stress in the material may be taken as heretofore, provided a 
satisfactory determination of stresses can be made. If this is not possible we must 
fall back on experience and judgment. 

1 This computation assumes a maximum fiber stress of 125000 lbs. per sq.in , but the maximum load, 
from the diagram, is only 165 instead of 185 lbs., so that the real maximum fiber stress is only 

X125000 = 111000 lbs. See note accompanying Table 39. 








GENERAL ENGINE PARTS 


205 


The drawings in Part III give a general idea of the valve gearing usually employed 
for internal-combustion engines. They show that in horizontal engines the valves are 
generally actuated by cams or eccentrics situated on a lay shaft, which is driven at 
half speed by gearing from the crank shaft and passes usually along what may be 
considered the front side of the engine. The valve gearing of vertical machines does 
not show such standard type of design. In the smaller size of this machine, with 
exhaust valves at the side of the combustion chamber and automatic inlet valves, it is 
comparatively easy to actuate the exhaust valves by the use of a single lay shaft 
parallel to the crank-shaft. If, on the other hand, the valves are placed in the 
cylinder head, the construction grows 
a little more complicated, since the use 
of a single vertical lay shaft does not 
give a good chance for a generally satis¬ 
factory design. In such cases an inter¬ 
mediate shaft transmitting the motion 
from the main shaft to the gear shaft 
horizontally mounted above is usually 
required. This intermediate shaft with 
its two gears (Figs. 258 and 259) makes 
the engine cost more, although compared 
with the generally very low cost of 
construction of vertical machines, the 
increase caused thereby is of small im¬ 
portance. The construction shown in 
Figs. 261 and 262 meets all require¬ 
ments for vertical engines up to say 20 H.P. The annular gear used works very 
quietly and accurately, and on account of its large contact surface is very well suited 
for high speeds. 

Based on experience it may be said that, for the transmission of motion from the 
main to the lay shaft, a pair of spur or screw gears is best and most satisfactory. 
Many attempts have been made, because of lower cost, to use for this purpose various 
types of clutches, chain drives, etc., but none of the substitutes so far have shown any 
considerable measure of success. Attempts have also been made to replace the cams 
so commonly employed for actuating the valves by other devices of equal value. 
Cams do the work required of them very satisfactorily. The only other construction 
to be considered would be the eccentric. Eccentrics in themselves, however, are not 
very well suited to gas-engine requirements, and when used in the earlier types of 
small engines have been able to maintain themselves only in isolated instances. They 
cost more than cams and are less simple, offer less range of adjustment as far as 
the motion of the valve is concerned and possess especially the disadvantage that 
the lift of the valve is comparatively slow. To avoid the latter drawback, either 
large eccentricity, which means large eccentric dimensions, must be used, or recourse 
must be had to complex valve gearing with its accompanying lost motion. Fig. 255 
shows that in such a case only about one third of the motion of the eccentric is 
utilized. 

It cannot be denied, of course, that cams also have serious disadvantages when 
used in engines above a certain capacity. The pressures between cam and roller at 
the beginning and end of the motion, and their reaction on the rest of the valve gear, 























206 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


finally grows very heavy, which becomes manifest in noisy operation, rapid wear of 
contact surfaces, etc. This fact has, in the design of large engines, led to the adoption 
of the eccentric valve drive in combination with a peculiar system of rocking valve 
levers by which the closing of the valve is mechanically controlled. (See p. 210, 
Figs. 278-282.) Other manufacturers, imitating the well-known valve gear construction 
of steam engines, have equipped their large engines with simple eccentric motions and 
attempted to get a satisfactory valve action, in combination with so-called wiper cams 
or similar devices. Whether these innovations will lead to any general use of the 
eccentric in the design of large engines will depend largely upon the practical results 
obtained. Information on this point is as yet quite meager. The satisfactory service 
of this type of valve in steam-engine practice can hardly be accepted as a criterion 
because the conditions of operation there are quite different (almost completely 
balanced valves, considerably smaller lifts, shorter time of valve action, etc.). 


Designs of Valve Gearing. 



Fig. 260. —Combined 4-cycle Valve 

Gear and Pendulum Governor, G. 

Wenzel. 

(Slide a and pick-blade b are ope¬ 
rated from the crank shaft. Sliding 
piece e with deflector d is held to 
e' only by friction and follows the 
motion of e' to the lowest position. 
In this position, d will deflect the 
ascending blade b, so that it will 
miss the stem e', but will strike e, 
moving this upward along e'. At 
the next upstroke however b will 
strike e' before it encounters d. If 
the speed is above normal, b will be 
constantly thrown out a sufficient 
amount to miss e' by the deflector 
c.) 



Figs. 261 and 262.—Valve Gear for an 8 H.P. Guldner Engine. 

(The spur gear a drives the annular gear b, which, on its cir¬ 
cumference, carries the inlet, exhaust and compression-relief 
cams E, A, and K .) 



Figs. 263-265.—Inlet Valve Gear Employing Eccentric, 6 H.P. 
Engine, Krupp-Grusonwerk. (Exhaust is operated by cam.) 








































































(Eccentric a , whose motion is guided through lever b, 
rotates the small gear c by means of the screw threads 
j'. Cam e, on the same shaft with c, operates the slide 
/, so that every other turn of the engine the blade g 
strikes the exhaust valve gear h.) 



Fig. 269. —Valve Gear, Old Type, Gasmotoren- 
fabrik, Deutz. 

(a, inlet’ b, gas; c, exhaust valve; d, open hot tube 
igniter. The valve levers are lettered the same as their 
respective valves.) 



Figs. 270 and 271.—Valve Gear, Motorenfabrik Werdau. 


(a, inlet; b, gas; c, exhaust valve; d, electric igniter; d', make-and-break cam; e, igniter; /, lever for regulation of airjsupply.) 





Figs. 272 and 273.—One of the 
Older Types of Nurnberg 
Valve Gear. 


(a, inlet valve; b, gas valve; c, exhaust valve; d, blowing-out valve; 
e, governor lever; /, slide block for gas valve rod; /', starting valve; a and 
V • ignition apparatus and igniter. The governor controls the position of 
the gas inlet rod b' along the block /. and thus determines the opening of 
this valve.) Angles a and are taken so that one cam operates both inlet 
and exhaust. 

















































































































































































































































































































































































208 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 










T 







'-Hr 




1 



lb 


Figs. 274 and 275.^Valve Gear used on some of the older Type of 1500-2000 H.P. Niirnberg Tandem Engines. 

(o inlet valves; b, gas valve; c exhaust valves; d, blowing-out valve; e, governor linkage;/, starting valve" 
g, valve lever springs. These engines are governed by adjusting the port openings through gas valve 6 the 
governor through e changing the fulcrum about which lever b" turns, and thus controlling the valve lfft 















































































































































































































































































































































GENERAL ENGINE PARTS 


209 



Fig. 276.—Valve Gear of a 600 H.P. Twin Engine, L. Soest & Co., Diisseldorf-Reisholz. 

(a, inlet valves; b, exhaust valve; c, gas governing valve; d, starting valve; e and /, double electric 
ignition; g, spark control. See Plate XXIII, Part III.) 



Fig 277_Valve Gear for a 100 H.P. Double-acting 4-cycle Engine, Langen & Wolf, Vienna. 

(a, inlet; b, ^gas; d, exhaust valve; e, igniter. The corresponding^operatmg Ifg™ ^ S 

same letters primed. Lever b', fastened to a , is thrown to the e > e. the left of b.) 

governor, through rod c changes the position of the guide block c, i 1 ° 













































































































































































































































a> 

> 


G 

> 


<N 

00 

<N 

I 


0> 

44 

cd 

o 

Q 

X 

o> 


II 


Oh 

a 


•♦o 

CO 

s 

« 

r< 

8 


a^ 

Q-> 

•» rn 


£ m 
N c 

. c3 
OQ ® 


<D 

G 


bX) 

G 


o O W 




73 

■*r c 

s tH 

<*> 

^ O 

S’C 

G 

o 
^ o 


OQ 

o 

Sh "3 

■ a> — 

<X> G 

^ G 


O O ^ £ 

43 43 $ g 
+3 o G 

>• o 
,73^ CD 

a ^ 

G • o 

7S {-, 


73 

I 




0) 


OQ 

G 

<D> 

CL 

O 


43 

to 

p 

o 

u 

43 

+-> 

0) 

> 


5 *p 

g .a 

CL U 
. Gh 
CQ 


■+^ 
I-* OQ 

c G 
G > O 
ri ox 
o —• ^ 


£ £» 


<D .2 


bC^j 


Sa°y 


"cc .£ G T3 


2 x 

f^-> £ 

+* G 

"N 


■s o * 

•§.2 £ 

G - « *> 
O^ 


co_L 

tc 73 
,rH <d 


§§ 


4^71 i ta 
0> 3 73—' 
73 S 94 ^ 


w 


-G 4? 

<d 

O.ES 


2 G 
G G3 


£ ~ 
V © 


43 ^ f—I ^ 
X M 

O OG -p 


o 

s-. 

G 


.y O 

cr O •—« 

oj ^ 

<D ^ 0 43 

f W 4 J 

o ^ *^ - 


a? h 

§.s 
~ H ^ 

oq a; 


—5 G *G 
G G g £ 

<j O ° 
OQ O ^ 

G £ c * rr 

bx)« a 

S G 


bX) <D^O 
G 43 

4-> 


ow s ^ a 

CD °0^ 


73 
G 1 
<D 
Qh 
<D 
73 


*^0^3 


• i—< >r 

^ o 




o 


G a> 
£ 43 

o^ 

<D r 

4G 


S ^.s 

O 43 
Jr V ^ 

^ rj ^ 
73 
<D 


a 
r-< oq 


0Q 

c3 

a> 


a> 


G -G 

o o 

p£ 


£ g « 
2't a 

C3 o3 G 

: j.° 

'o^SC 

-M . O 
G £*■> • 
-^4D4^ 

_G 73 O 

C_i ^ G 

5 8- 


i 

Sh 

<u 

<v 

73 


0Q 


t- G 1 

SS.gs 

% g>g-s 

“^.9- 

s ® £5 b 

•25 & £ 

*3.73 O.tJ 

a^ 1 ^ rt 

rG G £ 

073 SP^2 

OQ G ^ 


H) D 1) 
43 43 43 


^ G 
O OJ 


| 4J ^3 

-£"« es to 
2_,J3 2, 
S-o +> g 


I 

cc 

G 


bX)^ . 


ac 0 ^ 
O G -m X 


S- hh 

Jh 
G3 HH 


G 


S3 G OQ , 

go a? a> 
bX)£ 
C3 


<D 

43 

-^» 


&Q> 


G G 


> 0 ) 

SP* G ^ 

^43 
o 


1 S- 

G ^ 

Ph 
o ^ 

a £ 

\W-~ 




^.2o 


o 

s- 


0Q 


G 73 
bX)^ 


OQ 

. G 
0) o 
> +2 
•3 g 
> 5 










































































































































































































































































































































































GENERAL ENGINE PARTS 


211 


Constructive Details. 


Table 19 


DIMENSIONS OF TEETH OF SCREW GEARS 


Diametral 

Pitch. 

Circular 

Pitch. 

Depth of 
Tooth. 




Lateral Pitc 


tn 




“ I'S 

cos a 

a = 26° 35' 

a = 

30° 

a = 

45° 

a = 

60° 

a = 63° 25' 

=—= M 

= tn 

Total 

Face 

ts 

ts 

ts 

ts 

ts 

ts 

ts 

ts 

ts 

ts 

7r 





TZ 


7Z 


TZ 


TZ 


TZ 


Inches 

Inches 

Inches 

Inches 

Inches 

Inches 

Inches 

Inches 

Inches 

Inches 

Inches 

Inches 

Inches 

.08 

.251 

.173 

.08 

.280 

.089 

.289 

.091 

.356 

.113 

.502 

.16 

.560 

.178 

.10 

.314 

.216 

.10 

.351 

.112 

.360 

.114 

.444 

.141 

.628 

.20 

.700 

.223 

.12 

.377 

.260 

.12 

.421 

.134 

.434 

.138 

.533 

.170 

.754 

.24 

.841 

.268 

.16 

.503 

.347 

.16 

.563 

.179 

.580 

.184 

.713 

.227 

1.006 

.32 

1.124 

.358 

.20 

.628 

.432 

.20 

.702 

.224 

.725 

.231 

.890 

.283 

1.256 

.40 

1.404 

.447 

.24 

.754 

.521 

.24 

.842 

.268 

.869 

.277 

1.066 

.339 

1.508 

.48 

1.684 

.537 

.28 

.880 

.607 

.28 

.983 

.313 

1.016 

.323 

1.244 

.396 

1.760 

.56 

1.966 

.625 

.32 

1.006 

.693 

.32 

1.125 

.358 

1.219 

.388 

1.422 

.453 

2.012 

.64 

2.249 

.717 

.36 

1.131 

.780 

.36 

1.263 

.402 

1.305 

.415 

1.599 

.508 

2.262 

.72 

2.528 

.806 

.40 

1.256 

.864 

.40 

1.405 

.447 

1.449 

.461 

1.774 

.562 

2.512 

.80 

2.807 

.892 

.44 

1.362 

.953 

.44 

1.544 

.491 

1.595 

.507 

1.954 

.622 

2.764 

.88 

3.088 

.982 

.48 

1.508 

1.039 

.48 

1.685 

.537 

1.741 

.554 

2.132 

.678 

3.016 

.96 

3.371 

1.072 

.52 

1.635 

1.127 

.52 

1.823 

.580 

1.885 

.599 

2.308 

.734 

3.266 

1.04 

3.650 

1.160 

.56 

1.759 

1.213 

.56 

1.964 

.625 

2.030 

.646 

2.485 

.791 

3.518 

1.12 

3.932 

1.250 

.60 

1.885 

1.296 

.60 

2.108 

.671 

2.176 

.692 

2.665 

.847 

3.770 

1.20 

4.215 

1.340 

.65 

2.042 

1.408 

.65 

2.280 

.726 

2.358 

.750 

2.886 

.918 

4.084 

1.30 

4.565 

1.452 

.70 

2.198 

1.516 

.70 

2.456 

.781 

2.537 

.806 

3.107 

.988 

4.396 

1.40 

4.913 

1.563 

.75 

2.356 

1.625 

.75 

2.634 

.838 

2.719 

.865 

3.330 

1.058 

4.712 

1.50 

5.266 

1.672 

.80 

2.512 

1.732 

.80 

2.808 

.893 

2.900 

.923 

3.551 

1.128 

5.024 

1.60 

5.616 

1.787 

.85 

2.670 

1.842 

.85 

2.984 

.948 

3.083 

.981 

3.774 

1.200 

5.340 

1.70 

5.969 

1.896 

.90 

2.826 

1.949 

.90 

3.158 

1.006 

3.263 

1.038 

3.995 

1.271 

5.652 

1.80 

6.317 

2.017 

.95 

2.984 

2.057 

.95 

3.333 

1.060 

3.445 

1.097 

4.218 

1.340 

5.968 

1.90 

6.670 

2.123 

1.00 

3.142 

2.167 

1.00 

3.511 

1.117 

3.628 

1.155 

4.445 

1.414 

6.284 

2.00 

7.024 

2.236 


1. Driving Gears. The lay shaft is usually operated by means of cylindrical 
screw gears, of which the driver, on the crank-shaft, is generally steel, the driven gear 
may be bronze or cast iron. In case the width of the teeth is 


ample and oil-bath lubrication is employed, both gears are some¬ 
times milled out of cast iron. On account of the comparatively 
small surface of contact and the sliding friction always present in 
this type of gear, constant lubrication is practically a necessity. 
In order to simplify the construction of the bearings, it is desir¬ 
able to have the same diameter of pitch circle for both gears, 
or at least to make the diameter of the driven gear as small as 
possible, in spite of the fact that the speed ratio is 2:1. For 
that reason a common value for the angle a (Fig. 28o) is 63° 25', 
or 60° in round numbers, for the driving gear, and a =26° 35', 
or approximately 30°, for the driven gear. The lateral pitch t Sf 
if tn is the normal pitch (see Fig. 283), will then be 



ts — 


Dot 


tn 

cos a 


z 


( 1 ) 













































212 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 
This for the driving gear becomes, when 

a =63° 25', or a =60°, 

4= :h73 = 2 - 236 

and for the driven gear we will have, when 

a = 26° 35', or a = 30°, 

fc= lii6 =1 ' 118 ‘-SShT 1 - 155 *- 


The pressure on the teeth of the driven gear causes a thrust on the lay shaft. 
This is less with the smaller value of a for this gear. 

Table 19, page 211, gives some of the main proportions of screw gears. 

The width of the teeth is from 2-2.5 tn, the thickness at the pitch circle about .5 t n , 
that is, there is no back-lash. For practical values of tn see Table 20. Putting the 

diametral pitch — = M, we may also write 


Total depth of tooth = 2 \M, depth above pitch line, i.e., face of tooth, h' = M ; 
Pitch diameter 7)< = —, where Z = number of teeth; 

External diameter of gear, D a = A + 2— = Dt + 2M. 

7Z 


Table 20 


AVERAGE DIMENSIONS OF VALVE-GEAR PARTS 


Normal Pitch of Screw Gears = tn, Dia. of Lay Shaft = <-4, Dia. of Cam Rollers=d r . 


Nominal B.H.P. 

2 

5 

l(i 

15 

20 

30 

40 

50 

60 

75 

100 

of Engine. 












t n inches. 

.648 

.680 

.680 

.711 

.711 

.742 

.742 

.804 

.804 

.866 

.926 

d , “ 

1.25 

1.37 

1.37 

1.62 

1.62 

1.75 

1.75 

2.00 

2.00 

2.25 

2.50 

. 

d r “ . 

1.62 

1.75 

2.00 

2.25 

2.37 

2.62 

2.75 

3.00 

3.25 

3.50 

4.00 


2. Lay Shaft. The computation based upon the external forces acting upon the 
cams gives dimensions much too small for both lay shaft and transmission gears. 
Table 20 gives some practical values for these dimensions for engines up to 100 B.H.P. 
If the lay shaft is also called upon to actuate devices out of the ordinary, as, for 
instance, circulating pumps for cooling water or air compressors, the values for t n and 
d s in the table should be correspondingly increased, or, instead of screw gears, milled 
helicoidal gears should be used. These possess the advantage over screw gears in that 
the pressure is distributed over the surface of the flank of the tooth instead of bearing 
in one line only. In general, however, the lay shaft should not be called upon to do 

























GENERAL ENGINE PARTS 


213 


such work, as it reacts unfavorably upon the governor, especially if the power required 
is of fluctuating quantity. Even in ordinary operation it is often noticed in large 
machines that the varying torque to which the lay shaft is subjected during one 
revolution causes vibrations in the shaft itself, knocking in the gears and irregular 
governor action. The resisting torque in the lay shaft is a maximum at the time of 
opening of the exhaust valve. As the cam roller leaves the exhaust valve cam, the 
lay shaft receives an impulse through the action of the valve spring updn the valve 
lever. This causes a taking up of any back-lash in the driving gear if the torque so 
produced exceeds the frictional resistance of the shaft, which in large engines is nearly 
always the case. Several German designers therefore use a special fly-wheel on the 
cam shaft, which, by means of its inertia, prevents the reversal of the torque on the 
shaft, and hence contributes much to noiselessness of operation and steadiness of 
regulation. English and American designers very often eliminate the bad effect upon 
regulation of the varying torque in the lay shaft by the use of centrifugal governors in 
the fly-wheel, or by giving the governor an independent drive from the main shaft. 

3. Cams. Fig. 284 shows a complete cam system consisting of three cams, E for 
the inlet, A for the exhaust, and K to reduce compression at starting. Since the lay 



Fig. 284. —Normal or Standard 
Cam System. 



shaft, sometimes called the half-time shaft, has only one half the angular velocity of 
the main shaft, the values of the angles given in Fig. 284 must be doubled in order 
to refer to crank angles. The lines of rise and fall of the cams should be tangential 
to the base circle, and the transition from the tangents to the outer surface of the 
cam should be as gradual as the valve lift curve (curve h, Fig. 255, p. 201) permits. 
It is possible, by properly fixing the position of the cam-rollers, to actuate both the 
inlet and exhaust valve from the same cam (see Fig. 273). This simplifies the gear a 
little, but it is questionable whether the gain realized is sufficient to balance the 
attendant inaccuracies in the motion of the valves. Regarding the direction of motion 
of the cam shaft, it will generally be found by inspection that the lift of the valve 
levers is more easy in one direction than in the other. This of course guarantees less 
wear and noise. 

Large cam diameters permit of accurate adjustment of the valve motion, but 
knocking against the cam rollers is likely to result when the peripheral velocity of the 
cam approaches 3 ft. per second. Large cam diameters also increase the torque in the 
shaft. The method of fastening the cam to the shaft requires careful consideration. 
Cams movable in an axial direction have not proven reliable in large engines unless 
























214 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


the method of fastening is very carefully designed. An example of this is given in 
Figs. 285 and 286. 

The cam rollers are made of hardened steel. Common dimensions are given in 
Table 20. The width of the roller for the inlet cam may be made = .3 dr, that for the 
exhaust cam = .4 dr. These dimensions, however, should be checked to see that the 
maximum pressure does not exceed 3000 lbs. per inch of width of roller. 

4. Valve Levers. In order to keep the inertia forces as low as possible, the cross- 
section of the valve levers should be such as to combine greatest strength and stiffness, 
especially in the horizontal plane, with minimum weight. The return motion of the 
levers should be produced by special springs, and should not be left to the action of 
the valve springs. Suppose that the weight of the actuated parts of the valve gear is 
G lbs., referred to the center of gravity, and let the distance the latter is to be moved 
in t seconds be s, then the tension of the special return spring, also referred to the 
nenter of gravity as the point of application, should be at least 


Mx2s Gx2s Gsfi 

t 2 ** 32.2* 2 ^lfi* 2 


lbs. 


( 2 ) 


In this equation // is a correction factor which takes into account frictional resistances 
and unavoidable errors of computation. Practice has shown that it may be taken at 




Suction. 

r ^Starti ng 

h<Li/st 



fi= 1.25 to 1.50. 

The computations for strength of the 
valve levers, which hardly require any 
further explanation, should be based on a 


' -1-7 - - --- - - ~ 

Fig. 287.— Normal or Standard Valve Setting pressure at the moment of opening of at 

T^isvorrsvm QA ~ ~ _J n r 


Diagram. least 30 lbs. per sq.in. for the inlet and 75 

lbs. per sq.in. for the exhaust valve. It is 
possible, in engines using compressed air for starting, that through careless handling of 
the starting valve (taking air full stroke) the opening resistance of the valves may be 
in excess of the above figures. 

5. Adjustment of the Valve Gear. The valve gear may be quickly set with the 
aid of diagrams like Fig. 284 or 287. In either of these T pi and T pa designate 



Fig. 288. 


OC= 0 54 90 /26 /SO /<?6 90 Q° 54 &Q /26 /8Q /26 SO Sf 



7p a Tpi 7p a 

Fig. 289.—Valve Lift Diagram for a Twin Engine. 


respectively the inner and outer dead center position of the piston. To facilitate the 
operation of adjusting the valve gear in the shop it is well to fasten definitely only 
the gear on the main shaft and the cams, but to leave the gear on the lay shaft free 
until the final adjustment is made. To this end the seat of this gear may be turned 






































































GENERAL ENGINE PARTS 


215 


slightly conical, in the ratio, say of 1:12 or 1:10, and during the adjustment the gear 
is kept from slipping by forcing it upon the cone by means of a jam nut. After the 
.setting is completed and the engine finally tested out, the pair of gears should be 
marked as indicated in Fig. 288, and the driven gear may then be finally keyed in 
place. 

In designing the valve gears for multi-cylinder engines, a clear insight into the 
sequence of the various valve actions in all of the cylinders may be gained by drawing 
the valve lift curves in their proper phase with reference to one system of coordinates. 
In Fig. 289 this has been done for a two-cylinder machine. 


IX. Fly=wheels 

Material. In stationary engines usually cast iron throughout; in the largest sizes 
of engines sometimes cast iron with steel or cast-steel arms; for high-speed automobile 
engines now and then steel or cast-iron body with lead rim. 

Allowable Bending Stress Kb, on account of uncertainty of the nature of the 
stresses encountered and to guard against unforeseen contingencies, not to exceed 
3000 lbs. per sq.in. for cast iron or 7000 lbs. per sq.in. for soft steel. 

1. Determination of Weight of Fly-wheel. In theoretical investigations it is usual 
to employ the tangential effort diagram obtained from an assumed or real indicator 
diagram, to compute the moment GD 2 S or the moment of inertia of the rim. 

r = GD 2 s GD 2 S = GR 2 

4<7 ~ 128 g ' . (1) 

which is required for any given coefficient of regulation d. This method, however, is 
quite complicated and takes time. For that reason it is not much used in the drafting 
room, but since it gives a clear insight into the pressure and velocity changes it will 
be explained in detail. This will be followed by the development of a simple and 
purely mathematical method of equal practical utility. 

Plates II, III, and IV give a number of accurate tangential effort diagrams for 
single-cylinder engines; Plates V and VI similar diagrams for the common multi-cylinder 
combinations. All of these were constructed from actual indicator diagrams, 1 for 
the weight Gt or G 0 of the reciprocating parts, noted on the plates, and for connecting- 
rod lengths equal to L = 5r. All diagrams are drawn for the full operation, that 
is, for two revolutions for the 4-cycle, and for one revolution for the 2-cycle. In the 
case of multi-cylinder engines it is usually sufficient to draw the diagram for the 

interval between two successive ignitions. To construct these diagrams the first opera¬ 
tion is to redraw the lines of the indicator diagram, placing the various events side 

by side in a row, and spacing off the ordinates above or below the zero pressure 
line as required. Next draw the diagrams for the inertia forces according to page 156, 
and lay them in over the lines of the developed indicator diagram in proper order. 
The combination of the indicated pressures with the inertia forces then gives the 

diagram of net piston pressures. These in turn are used in the determination of the 
tangential effort diagram, which has for its length of base the developed crank circles 
of a complete cycle. The abscissae for the ordinates on all the diagrams are equivalent 


1 Published in the Zeitschrift d. V. D. I., 1901, p. 369. 



216 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


to 10 ° crank angle, and corresponding piston and crank positions are marked by the 
same figures. 

To determine the ordinates of the tangential effort diagram any of the well known 
graphical methods may be employed. Mathematically, their height is 


.p swft ibs . 


Vt = V 


cos /? 


per sq.m., 


( 2 ) 


in which p is the net piston pressure at any given instant, a=the crank angle, and 
/? the connecting-rod angle. 

To obtain the resistance diagram, we may assume that the net effective effort 
resulting from the fluctuating effort of one complete cycle is used up by a uniform 
resistance W acting tangentially at the crank circle. This diagram therefore is a 
fhctangle whose base b is the length of the developed crank circles of one complete 
cycle, and whose height p w may be determined from the mean effective pressure pi 
corresponding to the nominal indicated horse-power, to which the work of resistance 
must evidently be equal. 

The net effective work Ai is the difference between the total indicated work A a 
and the work of compression A c , and is produced during one stroke of the piston. 
Hence 

Ai=A a — Ac = viF X 2 r ft.-lbs.( 3 ) 

The uniform resistance encountered is 


W=p w Fb ft.-lbs.(4) 

The length b depends upon the type of engine as follows: 

For single-acting 4-cycle engines, b = 47zr; 

For single-acting 2-cycle engines, b = 2Tir ; 

For double-acting 2-cycle engines, b=izr. 

Since Ai=W or piFX2r=p w Fb, .( 3 a ) 

we obtain, by substituting the proper values of b, the following values of p w for all 
single-cylinder engines. 


4-cycle engines, .( 5 ) 

Single-acting 2-cycle engines, p w =—;.( 5 a ) 


Double-acting 2-cycle engines, = . 


The crank travel b between the successive ignitions is therefore determined in the 
various cases by the following crank angles: in the double-acting 2 -cycle, 180°; the 
single-acting 2-cycle, 360°, and the single-acting 4-cycle, 720°. In what follows' the 
pressures pi, p c , and p w will be expressed in pounds per square inch. If F is the 
effective piston area in square inches, the factors p,F, p c F, and p w F represent the 









PLATE II 





















































































































































































































































































































































PLATE III 



































































































































































































































































































































* 









































































































































































































PLATE IV 



/=•«• //o. S'1 77 . S^/.s/Z^f. /^/Z.es-spxT?. ^.TXr/^r 

/77Z. — /tPo/ZlS/ts^ir?. /lT7. ^SZOt&s/sp.-trr. 


_ - _ , Pw —<2. PP Lb&. 


Piston Pressure and Tangential Effort Diagrams for Single Cylinder Engines. 



































































































































GENERAL ENGINE PARTS 


217 


value of the total piston or resisting pressures Pi, P c , and P w acting at any given 
instant. 

The difference existing at any given instant between the variable tangential effort 
and the uniform resistance is shown in the diagrams by the relative positions of 
the two lines representing these forces. 

Where the tangential effort curve lies 
above the line of uniform resistance, we 
have an excess of energy furnished; simi¬ 
larly, where the curve cuts below the 
line of resistance, a deficiency exists. 

As is well known, this deficiency is 
made up by means of the kinetic en¬ 
ergy K derived from the energy excess 
A and stored in the revolving mass M 
of the fly-wheel. The energy exchange 
between crank and fly-wheel 
takes place under constantly 
varying angular velocity of the 
crank. The velocity diagram 
(Fig. 4, Plate II) gives some 
idea of the kind and magni¬ 
tude of the velocity varia¬ 
tions. It will be noted that 
is found at that point of the 



Mean \ 

Speed*] ___^ 

/ynition /fnit ion Jfnxtton Wss-Strote /<?nitt<m 

Fig. 290.—Tachogram, 100 H.P. Crossley Engine. 


the maximum velocity F max of the crank pin 
diagram (Fig. 3, Plate II) where the tangential 
effort cuts below the line of uniform resistance, that is, near the end of the 
expansion stroke. Beyond this stroke there is in general a steady drop in the 
velocity, interrupted only here and there by the effect of inertia forces of the recip¬ 
rocating parts, until the minimum velocity, F min of the pin is reached just after 
compression. After this point the tangential effort curve quickly rises again above the 

line of uniform resistance owing to the com¬ 
bined action of the inertia forces and the 
new combustion. 


Actual tachometer records, taken on a large 
scale with very sensitive tachometers, corrobo¬ 
rate throughout the statements above made in 
regard to velocity variation of crank pin as 
derived from the constructed tangential effort 
diagram. Fig. 290 shows a record for five com¬ 
plete cycles as obtained from a 100 H.P. single 
cylinder Crossley-Otto producer gas engine. This 
engine had two fly-wheels, each weighing 5 tons, 
and was supposed to make normally 180 turns 
per minute. During the tachometer tests the 
average r.p.m. was 164 at a load of about 60 H.P. 
The velocity variations shown in this diagram 
are remarkable for irregularity, magnitude, and 
unexpected occurrences. This is true not only as regards variation between cycles, but also 
within the individual cycles. It is evident that the fly-wheels were too light at the lower speed 
used and that the individual explosions were of very unequal strength. Much more regular, 
that is, uniform, are two records from a 20-H.P. Diesel engine, running normally at 180 r.p.m., 




Figs. 291 and 292.—Tachograms, 20 H.P. Diesel 
Engine. 



























































































218 


DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

shown in Figs. 291 and 292. The fly-wheel in this engine weighed about 3700 lbs. Fig. 291 is 
taken at 9.36 B.H.P. and 292 at 21.75 B.H.P. 1 


Assume that 


A = 


K = M 


V 2 a 


^ 2 min 

2 


ft.-lbs. 


( 6 ) 


If we let v= Vmax V — = the mean velocity of the crank pin and the coefficient of 
2 

regulation 

* V max ~ flmin n max ~ ^min .(7) 

V n ’ 

equation (6) may be written 

. . . . . . ( 6a ) 


Further, if the velocity V in ft. per second of the center of gravity of the cross- 
section of the wheel rim be substituted for v, and the weight G of the rim be put in 


place of M, we finally have 


A = K=-V 2 d. 
9 


( 8 ) 


Assume that the area representing energy excess or deficiency in the tangential 
effort diagram contains / sq.in. The energy value of this area will be 

A =fFmj ft.-lbs.( 8 °) 


In the above equation S = 2r=stroke in feet, i = scale length in inches of one stroke 
in the net piston pressure diagram, and m=number of pounds pressure per inch of ordi¬ 
nate of this diagram. 

The necessary weight G of the fly-wheel rim must be, from eq. (8), 



2900 A 
JizRn\ 2 ~d(Rn) 2 

d {so) 


From (9) we also derive 


Ag 2900 A 
GV 2 G(Rn) 2 


( 9 ) 


( 10 ) 


Under ordinary conditions the area of excess in the tangential effort diagram is 
equal to the area of deficiency, in which case the latter may be left out of considera- 


' These diagrams, and also those on page 227 and 228, were taken from an article by Prof A. Reeve 
in the Engineering News of Aug. 2, 1900. They were made by the use of an dectncally exeitrf 
tuning-fork! the stylus of which passed over smoked paper. The records so obtamed were afterwards 
carefully copied. 















GENERAL ENGINE PARTS 


219 


tion. It may happen, however, in engines governing on the hit-and-miss principle that 
that part of the area of the tangential effort diagram lying below the line of resistance 
exceeds the positive area of excess. In such a case the value of A must be determined 
from the deficiency area. This will be explained more in detail when we come to 
consider various methods of regulation. In Plates Y and VI, showing diagrams 
for 2-cycle and multi-cylinder engines, the areas upon which the computation of A 
should be based have in each case been indicated by marginal hatching. 

At the outset it may be stated that for 4 -cycle engines the inertia effects of the 
reciprocating parts may be neglected in fly-wheel computations. It is true, of course, 
that these inertia pressures tend to decrease the pressures on the crank pin at the 
moment of maximum compression and combustion pressures; it is also true that the 
inertia effects cause the tangential effort to be more uniform during the second and 
third strokes of the cycle; but all of these influences affect the coefficient of regulation 
8 of the engine but very little. The influence that the moving parts may have upon 
the area used as a basis of computation for the value of G or 8 is very insignificant. 
The difference usually amounts to but from .5 to .7%, and only in the most unfavorable 
cases does it amount to 1 % of the true area . 1 The explanation of this not quite 
self-evident fact is found in the circumstance that the line of uniform resistance, which 
forms the lower boundary of the excess energy area, lies very close to the line of zero 
pressure in 4-cycle engines. In fact, this line cuts the tangential effort curve of the 
expansion stroke close to the dead centers. In these piston positions, however, the 
crank angle is so small that the inertia forces do not produce any considerable 
turning effort, and the pure tangential pressure curve is therefore influenced but little. 
Since the positive inertia pressures during' any one stroke must equal the negative 
pressures, i.e., their sum = 0 , the absolute size of the work area is not affected by 
them. Their only effect is to change the areas above and below the line of uniform 
resistance by changing the shape of the circumscribing curves. But the amount of this 
change is the smaller the closer the line of resistance is to the zero line. In the case 
of 4-cycle engines the distance p w from the zero line to the uniform resistance line is 
so small that the designer is justified in neglecting the effect of the inertia forces in 
fly-wheel computations for engines of this type. (Of course the value of the correct 
tangential effort diagram in judging actual pressure conditions and reversal of pressure 
is not in any way affected by these considerations.) 

The mean pressure of the negative work done during the first and fourth strokes 
in drawing in the fresh charge and discharging the exhaust gases is only about 1 % of 
the mean pressure of the absolute positive work in modern large engines with mechan¬ 
ically operated valves of large cross-section. This work has therefore practically no 
effect upon the turning moments. In any case its effect upon the coefficient of 
regulation is less than that of the frictional resistance of piston and cross-head. But 
so little is definitely known about the magnitude and distribution of the latter over 
the four strokes of the cycle, that its effects cannot be taken into account with any 
accuracy in fly-wheel computation. For that reason we are justified in this investiga¬ 
tion in also neglecting the negative work of the suction and exhaust strokes, and 
confining our attention merely to the negative work A c of the compression stroke and 
the absolute positive work A a of the combustion and expansion stroke. 

The difference A a — A c represents the indicated effective work Ai, which, according 


Actual proof of this may be found in the Zeitschrift d. V. D. I., 1901, p. 371. 


220 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


to eq. (3a), must be equal to the uniform work of resistance W. In the tangential 
effort diagram of a 4-cycle machine (Fig. 293), A a is represented by the positive 
area abcde. After subtracting the area of the strip Ao=abde, the remainder is the 
area A = bed, from which is determined the rim weight G of the fly-wheel. The area 
Ao is smaller than the parallelogram aa\e\e only by the small corner areas x + y=aa\b 
+de\e. The area aa^e represents one quarter of the work of the uniform resistance 



The integration of a large number of turning effort diagrams of 4-cycle 


engines has conclusively shown that the area of the two small triangles x and y, cut 
from the parallelogram aa\e\e rarely exceeds 1%, and in the extreme does not amount 
to H% of the positive work area abcde. The influence of these areas upon the result 
of fly-wheel computations is therefore in all cases very small. 1 It is 'permissible, 
therefore, without introducing undue errors, to determine the excess energy A directly 

W W 

from the expression A a —Now both A a and can be taken directly from an 


indicator diagram, and the latter is consequently quite sufficient to determine the 

amount of excess energy A and finally also the 
weight G of the fly-wheel rim. This enables us 
to substitute a simple mathematical method 
for the graphical method. 

The term “absolute positive work” ( A a ) is 
not much used in actual practice. The usual 
method is to planimeter the indicator card and 
to determine from this the value of Ai, that is, 
the net indicated work, and for that reason it seems best in the method to be developed 
to use the more common terms Ai (in ft.-lbs.) and N{ (in H.P.). 

Let the mean pressure of the compression stroke be p c , and the mean effective 
pressure (of the net indicated work) be p% lbs. per sq.in. The absolute positive work 
will then be 



Now, putting 
we may write 

Excess energy 
But from eq. (3a) 
Hence 


for a 4-cycle machine. 


A a — + ft.-lbs. 


w w 

A=A a —^-=A»( 1 +p) ——ft.-lbs. 
Ai=W = p w Fb. 


A = W(l+p)-^^(.75+p)W 

= (.75 +p)p w Fb 
= (.75 +p)p w F 4 r 7i ft.-lbs. 


( 11 ) 

( 12 ) 


(13) 


1 For numerical proof, see Zeitschrift d. V. D. I., 1901, p. 371. 







GENERAL ENGINE PARTS 


221 


The uniform turning moment p w Fr = P w r, however, may be expressed by the 
general dynamic equation, 


D 550 N 60X550 N co _JV 

Pr— -= ~— —=5255—. 

co 2 nn n 


(14) 


Hence eq. (13), in connection with eq. (8), becomes 


(. 75 + p) 5255—4tt = A=—V 2 d, 
f n g 


(.75 +p)5255 NAng 2124600(.75 +p)Ni lbg (15) 

V 2 dn oV 2 n 


In round numbers we may therefore write 

g- 2125000 *; ( - 7s+ ^ ibs. 
oV 2 n 


(16) 


and 


2125000 Ni(.75+p) 
GV 2 n 


• (17) 


It is easier for both designer and constructor to use the radius R or the diameter 
D s of the gravity circle of the wheel rim instead of the velocity V in computation. 
Substituting therefore 


V = 


D s n 

“6(T = 


D s n 
: 19.10 


in eqs. (16) and (17), we finally have 


c 77.5 X 10 7 (.75 +p) Ni ^ 
dD 2 s n 3 


(16a) 


and 


77.5X10 7 (.7S >+p)Ni . (17a) 


GDin 3 


Of the total weight G about T V is usually assigned to the arms of the wheel, so 
that only G needs to be considered in designing the rim. Investigations on gas 
engines, however, have shown that the influence of the weight of the arms seems to 
be greater than the above proportion would lead us to expect, that is, the experi¬ 
mentally determined coefficient of regulation is smaller than a rim weight of G 
should theoretically give. The reason for this may be in the fact that the arms are 
usually designed much heavier than is common in steam-engine construction. 

It should be noted in eq. (17a) that the factor n enters in the third power, hence 
a comparatively small decrease in n increases the value of d seriously. On the other 
hand, the same facts permit us to materially better the regulation of any given 
machine by a comparatively minor increase in the number of revolutions per minute. 

The coefficient p, eq. (12), does not depend upon the constructive differences in 
various machines as much as it does upon the heating values of the fuel employed 
and the quality of the mixture. The purer and richer the mixture, the smaller will 

















222 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


be the value of p , because the relation between work of compression and of expansion 
is under such conditions the most favorable. With the high compression pressures used 
to-day in the best machines, however, one is compelled to use comparatively lean 
mixtures to prevent pre-ignition and its attendant trouble. These mixtures burn with 
comparatively small development of work area, and upon this depends the fact that 
in general the value of p varies directly as the compression pressure, rising and falling 
with the latter. 

To give some idea of the value of p, in approximate or preliminary computations 
for which p has not yet been definitely determined, the author investigated a con¬ 
siderable number of actual diagrams from the most important types of engines with a 
view to determining the values of A { and A c . Excluding the extreme cases, the 
following limiting values of p are found from the data so obtained: 


Table 21 


PRACTICAL VALUES OF p=— 

Pi 


Illuminating-gas engine. 


Gasolene engines. 

. ^ = .10 to .20 

Producer-gas engine. 


Alcohol engine.. . . . 

. ^ = .25 to .32 

Kerosene engine. 

. p = .30 to .40 

Diesel oil engine 1 . 



The value of Ni, in the case of engines governing down to no load on systems 
other than the hit-and-miss, should be based upon the intended maximum horse-power, 
N e max , of the engine, which should be about 10% greater than the nominal horse¬ 
power N n . If, as usual, the latter is determined at the brake, the mechanical 
efficiency r) m , which in the various constructed types varies from .70 to .85, must be 
taken into account. We may then write 

Vi=- max , 

T}m 


in which Ne max stands for the maximum brake horse-power desired. 

In engines using so-called “ precision ” regulation, the governing is effected by a 
change in the mean pressure of the combustion stroke. In the tangential effort 
diagram therefore the excess positive area decreases as the load decreases. The 


1 For these engines, Paul Meyer (Zeitschrift d. V. D. Ingenieure, 1901, p. 413) proposed the following 
special equation: 

„ 894.5 X.012 (A) J % 894.5 X .012(A) 

G =-kar. and 8 = -. 

dRW S GRV 

in which 2(A) represents the energy in m.-kg. absorbed by the fly-wheel, between the limiting velocities 
F ma j and Um n . This energy, for the ordinary type of Diesel engine, is approximately represented by 

2(A)=w(.77pi+4.1) cm.-kg., 

in which v is the stroke volume in ccm. 

[Note that in these equations the units are metric throughout.] 















GENERAL ENGINE PARTS 


223 


dislance, p-w, however, which measures the uniform resistance distributed over the four 
strokes of the cycle, decreases at a much slower rate. Hence the relation of excess 
area to resistance area grows more favorable as the load on the engine decreases. 
Consequently, also, the coefficient of regulation d, computed for N emax , constantly 
improves as the load falls. 

The conditions are much less favorable when the hit-and-miss system of regulation 
is used. Every decrease in the load in such an engine means an increase in the value 
of d, that is, less close regulation, a fact to be borne in mind when a certain value 
of d is to be maintained also at partial loads. The reduced diagram shown in Fig. 294 
serves to explain this. It is drawn for half load, i.e., theoretically there should be one 
miss between every two ignitions. The positive work A a developed during one com¬ 
bustion, considered as a constant turning force, is now distributed over eight strokes, 
instead of four as for full load, and the crank travel b 0 from ignition to ignition is 
2X4 nr. But since, as before, the indicated effective work Ai must be equal to the 
uniform work of resistance W = W 0 = p w0 Fb 0 , the value of p w0 must vary inversely as 
bo, i.e., it must be one half of its full load value. The line of uniform resistance is 


Miss-stroke 



iii / V 


\M 

r 

i 

t \f 

nr 

T 

IT |JJj|jr j 

JllfcJ 

1 W 


1 ■O'K i 




r'] 

■b a “2.‘f r X-■ - 

5-- 3 





Fig. 294. 


therefore that much nearer to the zero line, and the excess area A=abc of the full 
load is, at half load, increased to Ao = a\abcc\. We then have 


A 0 = A 


Wo 

2*4' 


(18} 


or, referring to eq. (13), 

A 0 =[(l — .125) -\-p]W 0 = (.875 +p) W 0 .(lga) 


The relation between the excess areas A and A 0 of the two load conditions, that is, 
in this case full load and half load, expresses directly the relation between the coeffi¬ 
cients of regulation d and d 0 . That is, 




.875+,0 - 
.750+/' 


(19} 


In the above discussion perfect governor action and ignition have been assumed; 
i.e., at half load, for example, one explosion must be regularly followed by one miss. 
If this is not the case, if, for instance, two, three, or more explosions are followed by 
a corresponding number of misses, the regulation will be still less close, since under 
these conditions the crank travel bo intervening between each governor action is- 
further increased, the distance pwo is further diminished, and excess area A 0 is still 
greater. 

If an engine, at any partial load Nio, works with x misses to one combustion, 
each of the latter must supply a crank travel equal to l+x = m full four-stroke cycles 



















224 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


or double revolutions. The coefficient of regulation then increases during the period of 
misses from the value o at full load to 


o 0 = 


1 


.250j 


m 




.750 +p 


( 20 ) 


If we assume an average value of p = .30, eq. (20) for various values of m, gives the 
following values of -y. 

m = 2 3 4 5 6 7 8 9 10 

| = 1.120 1.158 1.178 1.190 1.197 1.203 1.207 1.210 1.213 

(1.120) (1.150) (1.166) (1.177) (1.184) (1.190) (1.194) (1.197) (1.200) 

The values in parenthesis were obtained by measurement from a series of carefully 
constructed tangential effort diagrams, in which p and m correspond to the values 
assumed for eq. (20). The agreement in the two cases is evidently very satisfactory. 

The field of application of eq. (20), however, is limited. It applies only in cases 
for which the tangential effort diagram shows a positive excess area equal to or greater 
than the negative excess area for one governor action. 1 If this is not the case, if, for 
instance, the negative areas, the hatched area in Fig. 295, exceed the positive excess 



area, then the former determine the weight of the fly-wheel, or, in the case of engines 
already constructed, the coefficient of regulation. Eq. (20) then no longer applies. 
The positive and negative excess areas for one governor action are almost exactly 
equal at f load. In the full load diagram, Fig. 295, the positive area a is equal to 


'Translator’s Note. In this discussion the terms “governor action” and “governing period” 
are used. In explanation, suppose that an engine of the hit-and-miss type operates under a given 
fractional load b, and that with perfect operation of the governor, which is here assumed, the engine 
regularly gives z explosions followed by y misses. Then x + y = m is the “governing period,” i.e., the 
number of double revolutions or complete cycles passed through while the fly-wheel goes from minimum 
to maximum and back to minimum velocity. It is evident that the load on the engine may then be 
z 

expressed by b = — , and it is also evident that in one governing period the sum of the positive excess 
m 

areas must always equal the sum of the negative excess, or deficiency, areas, otherwise the engine will 
either stop or run away. 

By “ governor action,” on the other hand, is meant the period consisting of the last explosion 
stroke of a series plus the number of strokes concerned in the interval of miss cycles up to the 
beginning of the next explosion. It is evident that for one governor action the negative excess, or 
deficiency, areas may exceed the single positive excess area. 

, 111 1 „ 

In the case of the partial loads represented bv the series —, —, —, to —, where z = 1 in all 

. 2 3 4 m 

cases, the governing period is the same as the period of governor action. 
















GENERAL ENGINE PARTS 


225 


the sum of the negative areas b. If now the load drops to .9 Ni, for which—assuming 
perfect governor operation—there should be regularly 9 hits followed by 1 miss, the 
unifoim lesistance oidinate p w will decrease to ^ 9 0 - its full load value, and the positive 

excess aiea increases by ^ ^ , At the same time, however, the sum of the negative 

excess, oi deficiency, areas also increases because these areas are now distributed over 
7 instead of over only 3 strokes, as at full load. It follows that, as the negative 
aieas increase at the expense of the positive excess area, during one governor action, 
the regulation coefficient increases, although immediately after the period of retardation 
it is brought down again to its best value as long as ignition regularly follows ignition. 

Prof. R. Mollier, 1 investigating this question of hit-and-miss regulation along the 
lines above laid down, has developed the following special equation for the numerical 
determination of the variation of d with different loads on the engine (see also 
eq. (18a)). 

A 0 =Ai(\ -^+j i +3r). ■ ' (I) 

In this equation 

•A 0 = as before, the excess energy during the period of governor action; 

Ai = A a — A c = net work of expansion; 

A c = work of compression; 

m = double revolutions (complete cycles) during one governing period; 

b=— — load on the engine. 
m 

Equation (I) thus assumes that each explosion is regularly followed by a = m — 1 
miss cycles. If, however, instead of one there should be z ignitions in one governing 
z 

period, then b = — . 2 

vn 


1 See Zeitschrift d. Y. D. I., 1903, p. 1704. 

2 Translator’s Note. As an example, suppose an engine governs regularly | 

11 5 

five hits followed by two misses. Then m = 7, — = — and b = —. Hence, from eq. (I), 

m 7 7 

1 3 5 


—, i.e., 


. / 1 3 5 A c \ /39 A c \ 


For full load, on the other hand, i.e., when no misses occur and m=co, eq. (13), page 220, gives 

A f 


A ~v 5+ z) Ai 


A c 1 

Putting — = — in both equations, we have 
Ai 4 

for load 
and for full load 


A 0 =H Ai, 
A =Ai. 


Hence the excess area to be considered at 4 load is greater than at full load, and if the fly-wheel 
is designed for full load, the coefficient of regulation d at full load will increase at f load to 


A, 


£=— 8 = 145 = 1.645. 

A 1T 

This is indicated in Fig. 296. It should be noted next that the condition that there shall be 
no misses at full load is ideal, and never realized in practice in hit-and-miss engines. A miss cycle 





226 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

It is evident that the greatest safety against any change in d exists when in 
eq. (I) the value of m is made very large, or, what amounts to the same thing, 

— or — is made very small. With — = 0, eq. (I) takes the simpler form 
mm m 

A 0 = A i (l+lb+fy .(II) 

From this equation of “practical regulation coefficients” Mollier has given in the 
diagram, Fig. 296, for various assumed values of z (ignitions) and a (miss cycles), the 



d Ac 1 

ratios of ■*-, under the assumption that ~r=j in all cases. Line AB gives the ratios 
O o -At 4 

when 2 = 1, which is the most favorable case, while line BC for — = 0 (most unfavorable 


case), gives from eq. (II) the only values of practical use. Line BC is the line of 
reference not only for the condition that any small load fluctuations occur at con- 


will always now and then occur no matter what the interval. Hence eq. (I) above instead of eq. (13) 
should be used for computing A for any load no matter how little below the ideal full load. For 
full load eq. (I) gives 


A-Ai 



which, with =—, becomes A=2A{. This also is indicated in Fig. 296 by the point C, and shows 
Ai 4 

that as soon as the ideal condition of no-miss cycles at full load is departed from the original value 
of S is increased very materially. 















GENERAL ENGINE PARTS 


227 


siderable intervals, but also when these fluctuations, grouped about the mean value of 
load, occur at intervals of any desired length. For this reason at f load, for instance, 
the regulation coefficient may still be computed from eq. (II) if the governor acts as 
indicated in the series: |||-|HIHH|H|-. Thus not only the effects due to load 
fluctuations, but also those due to any lag in the governor, are to a certain extent 
taken care of. 

The line BC, Fig. 296, shows that the closeness of regulation in practice increases 
under decreasing loads. 


The tachometer record, Fig. 297, shows with special clearness the magnitude and suddenness 
of speed changes sometimes occurring in hit-and-miss engines. At the end of the expansion 
stroke of the second cycle the number of revolutions per minute is 3% above, and at the end 
of the fourth, which is a miss cycle, it is 3% below the normal speed. Inside of these cycles, 
therefore, the crank will show a speed change of 6%,—and this with a fly-wheel weight of 
22000 lbs., which in this case amounts to 220 lbs. for each nominal horse-power. It should be 
noted, however, that this record represents extraordinarily unfavorable operating conditions. The 
upper speed limit (103%) could only have been reached by a very heavy explosion, while the 



Fig. 297.—Tachogram, 100 H.P. Crossley Engine. 


lower limit is in part at least due to the partial failure of ignition in the previous cycles. 
Under less unfavorable circumstances, the speed limits are much narrower than this, as is shown 
by Fig. 298, which is obtained from the same engine under the same load as for Fig. 296. If 
the power of an engine is transmitted by belting, the speed variations, owing to the equalizing 
action of the belt, will be much less pronounced in the power consumer, especially if this itself 
possesses rotating masses. Under such conditions there will be established in the power consumer 
an average speed for every four-stroke cycle, which is much closer to the normal speed than that 
measured in the prime mover itself. The advantage of equalization possessed by this elastic 
transmission is, however, lost if the speed increase or decrease in the engine extends over a 
considerable number of revolutions. This is clearly shown in the record of Fig. 298, for which 
the 100 H.P. engine was loaded only with the friction of a transmission about 100 ft. long. 
Here from 4-5 miss cycles are followed by 2 or 3 ignitions. During the former the average 
speed slowly decreases, while during the latter it increases by rapid stages. The maximum and 
average variations occurring under these conditions are well shown. 


While it is true that speed fluctuations of the kind outlined do not seriously 
affect operation in ordinary industrial applications, they badly interfere with satis¬ 
factory electric lighting service and other operation requiring close regulation. To be 
on the safe side in cases of that kind, the designer has only one recourse: to make 

the fly-wheel, in the case of hit-and-miss regulation, at the outset twice as heavy as 

the computation (assuming a given value of d for full load, i.e., without any miss 

cycles), seems to call for. Besides this, however, the greatest care should be taken not 

only in the design of the governing mechanism but also in the features of the mixing 
and ignition apparatus, in order to restrict the number of consecutive miss cycles to as 
small a number as possible and to insure that the first explosion following a period of 
retardation shall occur at the proper moment and be of full strength. But even under 












































228 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 



the best conditions, hit-and-miss regulation will require compara¬ 
tively large rotating masses in order to satisfy the closer degrees 
of regulation demanded to-day at all loads on the engine. 

As may be seen from the above discussion, the method of 
fly-wheel computation developed depends entirely upon the assumption 
that the small triangular areas x -F y, which are a part of one quarter 
of the total uniform resistance (see Fig. 293) may be neglected in 
comparison with the positive work area A a . In the case of the 
single cylinder 4-cycle and the old 6-cycle, this assumption is justi¬ 
fied; but the case is different in engines in which the crank travel 
between successive ignitions is shorter. With any given engine power, 
Ni, the ordinate p w of the uniform resistance varies inversely with 
the length b of the crank travel. The shorter this travel, the 
greater this ordinate [see eqs. (5) to (56)], and owing to this fact 
x ~\~y 

the area- ratio —~~ increases very rapidly, almost as the square. 


Thus, while in the single-cylinder 4-cycle (/? = 720°) this ratio 
does not usually reach .01, in the 2-cycle and the tandem or 
double 4-cycle (/? = 360°) it has already increased to from .03 to 
.04, and in the case of the double-acting 2-cycle or the double¬ 
acting two-cylinder 4-cycle (/?=180°) it even reaches .15 to .16. 
Such quantities can of course be no longer neglected, and in order 
to take them into account eqs. (16), (17), or (16a) and (17a), require 
the following modification: 

If, for example, a 4-cycle engine with a wheel rim weight of 
G lbs. is operated as a 2-cycle, the crank travel /? between ignitions 
decreases from 720° to 360°, and hence a smaller weight of rim, 
equal to, say xG lbs., is sufficient to maintain the former degree 
of regulation. Further, making the engine operate as a 2-cycle 
doubles the previous capacity of the 4-cycle machine, so that for 
the same d and the same Ni as before a rim weight of only 
.5 xG = kG is really required. The factor x or k has very different 
values depending upon the length of crank travel, i.e., upon the 
number of revolutions between ignitions, but is, on the other 
hand, little influenced by the shape of the indicator diagram, as 
was shown from a number of tangential effort diagrams. This 
fact makes it possible to regard k as a constant for the same 
operating cycles and similar cylinder or crank combinations. Eqs. 
(16) to (17a) then take the following form: 


G = 


k 2125000 Ni(.75+p) 
dV 2 n 


lbs.; 


( 21 ) 


d — 


k 2125000 Ni(.75+ P ) 
GV 2 n ’ 


( 22 ) 


or 


k 77. 5X10 7 {.75+p)Ni ^ 
dD 2 s n 3 

k 77.5X 10 7 (.75+ < o)At 
GD 2 n 3 


. (21a) 


. (22a) 










































































PLATE \ 



Partial Tangential Effort Diagrams for Multicylinder Engines. 















































































































































PLATE VI 




Partial Tangential Effort Diagrams for Multicylinder Engines. 








































































































































































































































































































































































































PLATE VII 













































































































































































































































- - -X 






> - -% — . ' 

- 























- 




























. 






































































GENERAL ENGINE PARTS 


229 


For values of k for the usual types of engines and operating cycles, see Table 22. 
The figures there given were determined from the tangential effort diagrams of 
Plates V, VI, and VII. In the construction of these diagrams, Figs. 5, 6, and 7 
(Plate II), referring to a Korting engine for which Z> = 14.95", S = 26.2", and n = 
140 r.p.m., were used as the basis. The turning effort diagrams of Plates V, VI, and 
VII therefore strictly apply only to one engine for which the indicator diagram showed 
p — 3 3, but it is permissible, from what has been shown, to regard the values of k in 
columns 5, 6, 10, and 11 of Table 22 as sufficiently accurate for all practical compu¬ 
tations. For comparing the respective values of the coefficient k the 4-cycle machine 
has been considered as the standard of reference. The values of k for the other 
types and combinations show directly how much more favorable are the governing 
conditions in these cases. Columns 6 and 11, giving the values of k 2 and * 4 referred 
to the same engine capacity Ni, are the best to use in making direct comparisons, 
especially if it is desired to determine the saving in fly-wheel weight due to a given 
combination of cylinders. 

The disadvantage of the single-acting single-cylinder 4-cycle engine with respect to 
fly-wheel weight necessary for a given engine capacity and coefficient of regulation is 
clearly apparent from columns 8 and 13 of the table. 

For this type of engine, with ^ = an d V = 65.8 ft. per second, a fly-wheel weight 
(including weight of arms) of 148.5 lbs. per I.H.P. is required. To satisfy the same 
requirements in a single-acting 2-cycle cylinder requires only 59.6 lbs. per I.H.P.; in a 
double 2-cycle, according to combination No. IV or VI (see table), it requires only 
12.5 lbs.; and in a three-cylinder 2-cycle only 5.94 lbs. per I.H.P. Compared to this 
last value, even the three-cylinder 4-cycle still requires 33.7 lbs. per I.H.P., that is, a 
fly-wheel weight about six times greater. 

With the aid of double action, the coefficient of regulation in the single-cylinder, 
single-acting 4-cycle improves in the ratio of 148.5 to 91.3 (see Table 22, column 8), 
that is, by about 38%. That in itself is considerable, but is only one half the amount 
gained by making the single-cylinder 2-cycle double-acting (from Table 22, column 13, 
59 6 —15 6 

the gain is = — ^ =74%). In the building of large gas engines at present the 

double-acting 4-cycle and the single- or double-acting 2-cycle are the types of engines 
most often met in competition. Between these the following relations exist with 
respect to the coefficient of regulation, as taken from Table 22: 


4-cycle. 2-cycle. 


Single Acting. 

Double Acting. 

Single Acting. 

Double Acting. 

1 000 

.615 

.401 

.106 

1.626 

1.000 

.652 

.172 

2.494 

1.533 

1 000 

.254 

9.434 

5.802 

3.783 

1 000 


Table 23 following, based on the preceding discussion, has been constructed for 
practical use. The quantities .9 Gi and GD 2 S were determined from eqs. (16) and (21) 
or (16a) and (21a), using the assumed average values of k, p, V , and n, as givfen. 
For any conditions not given in the table the desired results may easily be obtained 
directly from the proper equations. The constant C results from the contraction of the 
factors 


k 77.5X10 7 (.75+p) =C, 


(23) 




230 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Table 22 


REGULATION FOR VARIOUS CYLINDER COMBINATIONS, ASSUMING 7 = 65.8 Ft. per Sec., 

d = AND p=. 33. 


_ 

^3 

S 

3 

£ 


eg 

3 & 

a >, 

H 


ai 


e 

o 
O 

o 

I 5 £ 

*- 'o 

| t! 5 
COO 


>> 

o 




o 


o 

£ 


4 

5 

6 

7 

8 

9 

10 

11 

12 

13 

4-cycle Engines. 

2-cycle Engines. 

Crank Travel /? between 
Ignition in Degrees. 

For Equal Dimensions, 
D, S and n. 

For Equal Max. Capa- 
“ city, Ni max. 

K 

eJ 

s 

•<s> 

H.P. 

fa 

<3 S 

II 

6“ 

Lbs. 

Crank Travel /? between 

Ignition in Degrees. 

Same, D, S 

and n. 

o * 

& E 

S >> 

eg .ti 
m o 

K 

d 

s 

•«» 

H P. 

f * 

a 

cu s 
» 

II 

S3 

Lbs. 

as in 4-cycle 
Engine, which 
is taken =1.0 

«3 

*4 

720 

1.000 

1.000 

65.8 

148.5 

360 

.802 

.401 

131.6 

59.6 

540 180 

1.230 

.615 

131.6 

91.3 

180 

.424 

.106 

263.2 

15.6 

360 

.796 

.398 

131.6 

58.8 

360 

1.595 

.399 

263.2 

53.8 

540 180 

1.290 

.645 

131.6 

95.6 

180 

.335 

.084 

263.2 

12.5 

360 

.792 

.396 

131.6 

58.6 

360 

1.602 

.401 

263.2 

59.6 

540 180 

1.290 

.645 

131.6 

95.6 

o 

00 

t-H 

.335 

.084 

263.2 

12.5 

240 

.678 

.226 

1 

197.4 

33.7 

120 

.237 

.0395 

394.8 

5.94 

180 

.335 

.084 

263.2 

12.5 






180 

.335 

.084 

263.2 

12.5 








For the double-crank 2-cylinder opposed engine (/? = 360°) the values of combination No. V, and for 
the 4-crank, 4-cylinder opposed engine (/?=180°) the values of combinations Nos. VIII or IX are nearly 
exact. 

from eq. (16a) or (21a) assuming £ = .35, and taking the proper value of k from Table 
22. With this simplification the rim weight required for a given machine with p = . 35 
is then 

„ CNi u 

G “VrfTi‘ lbs .(24) 

oD s n 3 

and the moment is 

CN 

GDs= ^ni {t '~ .(24a) 

Hence 

d = Nj 

C GDZn* . 


(25) 























































GENERAL ENGINE PARTS 


231 


Equations similar to the above, involving a constant C, are well known in steam- 
engine design. The mean rim velocity V — 65.8 ft. per sec., upon which the results of 
Table 23 are based, is good for horse-powers up to about 15 or 20. V increases 
gradually with engine capacity to the limiting value of from 120 to 130 ft. per sec. 
(see p. 234). As will be seen from eq. (21), since V enters as the square , any varia¬ 
tion in its value will have a marked effect upon the value of G. Hence the importance 
of the proper choice of rim velocity is apparent. The set of curves on Plate VII, the 
use of which is there explained, will be found of great utility in this respect. 


Table 23 


REQUIRED FLY-WHEEL RIM WEIGHT AND MOMENT GD 2 S FOR F = 65.8 Feet per Sec., 
tc = 200 r.p.m., 8 = ^, p=. 35 AND k FROM TABLE 22. 


1 

2 

3 

4 

5 

6 

7 

8 

9 

10 




4-cycle 

Engines. 



2-cycle Engines. 


Number. 

Number of 
Cylinders and 
Cranks. 

Crank Travel /?, 
Degrees. 

Rim Weight .9 Gi 
required per 

I.H.P., Pounds. 

GD 2 S required per 
I.H.P., Ft.-lbs. 

Constant C from 
Eq. (23). 

Crank Travel /?, 
Degrees. 

Rim Weight .9 Gi 
required per 

I.H.P., Pounds. 

GD 2 S required per 

I.H.P., Ft.-lbs. 

Constant C from 

Eq. (23). 

I 

- 

Single-acting 

Types. 

1 cyl., 1 crank 

720 

133.5 

4265 

85.3 X10 7 

360 

53.7 

1710 

34.2 X10 7 

II 

2 cyls., 1 crank 

1540 180 

86.2 

2745 

54.9 X10 7 




7.16X10 7 

III 

2 cyls., 2 cranks 

360 

53.0 

1695 

33.9 X10 7 

180 

11.2 

358 

IV 

3 cyls., 3 cranks 

240 

30.1 

962 

19.25X10 7 





V 

4 cyls., 2 cranks 

180 

11.2 

358 

7.16X10 7 





VI 

4 cyls., 4 cranks 

180 

11.2 

358 

7.16X10 7 





VII 

Double-acting 

Types. 

1 cyl., 1 crank 

540 180 

82.3 

2620 

52.4 X10 7 

180 

14.2 

452 

9.04 X10 7 

VIII 

2 cyls., 1 crank 

180 

11.2 

358 

7.16X10 7 




1.27 X10 7 

IX 

2 cyls., 2 cranks 

180 

11.2 

358 

7.16X10 7 

90 

2.0 

68.5 

X 

4 cyls., 2 cranks 

90 

2.29 

71.5 

1.43 X10 7 



.... 



Under normal conditions of operation fly-wheels are commonly designed for the 


following coefficients of regulation: 

For ordinary power purposes and similar installations .... d = ^ to 
For electric lighting service (without storage batteries) with 

direct-current generators. 0 = 7 V to tit 

Alternating-current generators. to 


In case of operation of generators in parallel 1 and for direct connected generators, 
these values of d should be further decreased by from 30 to 60%. 


1 For thorough discussions of this question, see Zeitschrift d. V. D. I., 1904, p. 793, and Electro- 
technische Zeitschrift, 1902, No. 14. 







































232 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


2. Determination of Dimensions. 

Fly-wheel Designs. (See also the examples 
shown in the Plates of Part III.) 


Fig. 299.—Fly-wheel, 6 H.P. Banki 
Engine, Ganz & Co., Budapest. 

D= 5.5", 

5=10.0", 
n= 300 r.p m. 

(The projecting balance weight de¬ 
stroys the symmetry of the wheel and 
bothers while starting. The necessity 
for such weight should therefore be 
avoided.) 




Fig. 300. —Hub and Belt Pulley for Wheel shown in 
Fig. 299. 

(The purpose of dividing the hub is to avoid casting 
stresses in the arms. The joints are filled in and the 
shrink rings put into place before the hub is bored.) 


Fig. 301. —Fly-wheel, 25 H.P. Wenzel 
Engine, Fr. Zimmermann A.-G., 
Halle S. 

D= 13.4", 

5=18.9", 

n=160. 

(The cutting away of metal in that 
part of the rim on the same side with 
the crank serves to help balance the 
rotating parts. Note the strengthening 
of the arms where the belt pulley is 
attached.) 



















































































GENERAL ENGINE PARTS 


233 




Diameter of wheel 14.75 ft., face 24". Wheel 
is drawn into place by means of four 1J" screws 
which are held in a two-part ring let into a groove 
in the shaft. This is for the purpose of insuring 
proper seating and true running. 


Figs. 302-304. —Fly-wheel for a 400 H.P. Crossley Engine. 


On account of the unusual width of face, the arms are made section. Rim connection is made with 
5 bolts on the inside and 3 countersunk links (a) on the outside. 


Constructive Details. The static computation of fly-wheels and especially the 
check computation regarding the stress relations chosen is possible only on the basis 
of a number of more or less arbitrary assumptions, since neither the magnitude of the 
stresses occurring nor the method of their action and distribution in the wheel can be 
definitely determined. To be on the safe side, the most unfavorable set of conditions 
among those that may occur should be chosen, and the dimensions determined accord¬ 
ingly. Besides this, the allowable stress, especially in the arms, should be kept low. 
In this respect it should be noted that the stress in the arms, due to the constant 
energy exchange between wheel rim and crank, becomes a reversing stress, varying 
from a positive to a negative maximum value. 

The weight of a wheel rim having a cross-section of / sq.in. and a radius of R ft. 
to the center of gravity of the rim cross-section may be expressed by 


„ 2 Rnf 450 
G= 144 


= 19.63 Ay lbs., 


(26) 


in which 450 lbs. is the weight of cast iron per cubic foot. A given weight, G lbs., 
therefore requires a cross-sectional area in the rim of 

^ = l9^ffi Sq - in . (27) 

The diameter D s of fly-wheels, to the center of gravity of the rim cross-section, is 
usually taken at from 4 to 6 8, or R = 2 to 2.5 S. Large engines, however, sometimes 
compel the use of R = 3 to 3.5 S. The maximum value of D s depends upon the 
allowable value of the rim speed, which may be found from 


V 




31 gKz 


ft. per sec., 


T 


(28) 































































234 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


in which K z may be taken at 1700 lbs. per sq.in. for cast iron, and y= specific gravity 
of cast iron = 7.25. The allowable velocity therefore is 

„ J2.31X 32.2X1700 100rw , 

I max — \ rj —132.0 ft. per sec. 

This limiting velocity, however, should be approached oply in extreme cases, because 
the fly-wheel rim is not only put under a tensile stress equal to 

o z = . 09727 2 lbs. per sq.in.,.(29) 

as eq. (28) assumes, but it is subject also to considerable bending stresses due to the 
action of the arms. Hence the value 7 = 100-110 ft. per sec. should not usually be 
exceeded. 

The number of arms used in cast-iron wheels is usually six, unless the wheels 
exceed 10 ft. in diameter, when eight are employed. For widths of wheel face greater 
than 15 in., either arms of special cross-section or double 
spiders should be used. The dimensions of the arms depend, 
on the one hand, upon the bending moment due to the 
tangential force P (Fig. 306), and, on the other, on the 
tensile stress induced by the centrifugal force in the rim. 
The bending moment does not reach its maximum value 
owing to the pull of belt or rope, but owing to the inertia 
of the wheel rim should the engine be suddenly started or 
stopped. If we make the assumption 1 that at the moment 
of starting the maximum crank effort T max occurs when the 
wheel is standing still (7 = 0), a condition strongly improbable and with certainty always 
preventable, the inertia of the wheel rim will cause at the hub end of each one of 
the arms (i in number) a bending moment equal to 

M6 = ^- r Zin.-lbs.(30) 


P 



The bending stress then will be 

Mb Mb 10 Mb,, . /01 . 

i = Tr = .098Mi~'^56- lbs - persc l' ln -.< 31) 

in which a is the major and b the minor axis of the elliptical cross-section of the 
arms at the hub end. It is also here assumed that the radius of the center of gravity 
of the rim section belonging to each arm is R instead of R 0 . In reality i£ 0 = .94 to 
.96 R, that is, J? 0 is from 4 to 6% smaller than R. 


1 It would seem at first sight that the bending moment on the arms would be a maximum when, 
under full velocity (7 max ), owing to pre-ignition, the crank suddenly receives an impulse in a 
direction opposed to the direction of rotation. In actual operation this is not the case, since early 

ignitions with full development of explosion pressure (P 2 ) can only occur near the dead center. But 
the turning moment produced by them in these positions is always less than the bending moments 
due to the maximum tangential effort. 















GENERAL ENGINE PARTS 


235 


Too great tapering of the arms should be avoided; a good rule to follow is to 

decrease a by .18 R in., and b by .10 R in. for the rim cross-section of the arms, 

where R represents the radius of the wheel in feet. 

The centrifugal force due to the part of the rim belonging to each arm is 


MV 2 _GV 2 ^ .031 GV 2 _ .00034 GRn 2 
iR giR iR i 


. (32) 


This force tends to rupture the arm in tension at its weakest place. The resulting 

C 

tensile stress is o = -j- lbs. per sq.in., which for elliptical arms is 

Ja 

‘ , = 78L6 lbs - perSq - in . (33> 


The maximum tensile stress usually occurs at the outer ends of the arms. If, 
however, the arms should be weakened in any way, as for instance, by the drilling of 
holes for the bolts fastening the belt pulley, a check computation of Ob and a is to be 
carried through for these cross-sections. 1 

The connections of split wheels are subject mainly to the disruptive action of the 
centrifugal forces acting normally to the plane of division. It is important therefore 
that these fastenings take up this stress as directly as possible, i.e., without setting up 
any considerable bending moments. Since the load produced by these forces is never 
equally distributed over the rim and hub connections, and since there is no means of 
determining what this distribution is in any given case, the usual method is to make 
each of these connections strong enough to safely 
withstand by itself the centrifugal forces on one- 
half of the wheel. Shrink rings or similar means 
should not be depended upon to any extent 
because, even with the wheel at rest, they may 
be already stressed to the elastic limit. 

The sum of the centrifugal forces acting on 
the connections of a divided wheel consists of 
those ( Ck ) due to half the wheel rim and those 
(C 0 ) due to the arms of one half of the wheel. 

Both of these should be referred to the plane of division and considered as acting at 
their respective centers of gravity. Thus, with reference to eq. (32) and Figs. 307 and 308, 

C k = .00034 Rn 2 % X - = .00011 Rn 2 G lbs.,.(34) 

Z 71 



and for each arm of weight Ga, 

Ca = . 00034 G a r 0 n 2 sin a lbs. 


(35) 


1 Translator’s Note. In many cases designers also check the hub end of the areas for tension 

due to centrifugal force and add this tensile stress to the bending stress as found from eq. (31) to 
find the total maximum stress. 












236 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


For the middle arm sin a = sin 90° = 1.0. For the two side arms sin a = sin 30° = .5 
when i = 6, and sin a = sin 45° = .707, when {=8. The total centrifugal force due to the 
three arms will then be, referred to the plane of division, when i= 6, 


C a = .00034 G a r 0 n 2 (1 + 2 X .5) = .00068(?ar 0 n 2 lbs. 


(35a) 


and when i=8, 


C a = .00034 Gar 0 n 2 {\ + 2 X .707) = .00082 G a r 0 n 2 lbs.(356) 


The sum of the forces acting on the connections will therefore be 

2(C) =C*+C a lbs.(36) 

The method of computing the stresses existing in the connections will be shown 
by the following numerical example. This shows that under certain circumstances 
these stresses may be quite considerable, and in properly taking care of them the 
designer is limited by the fact that for large values of R or V the joint should be 
made as light as possible to avoid additional centrifugal stresses. If this cannot be 
readily done, it is better to cast the rim in one piece and divide the spider only. 

No computation is possible for the hub. Experience has shown that the following 
dimensions are satisfactory: 

Length of hub= 1.5 to 2.5 d (the latter value only in small engines) and external 
diameter = 2 to 2.5 d, where d = shaft diameter. 

It is usual to core out the middle of the hub for about one third of its length. 


Example. Computation for a split wheel for an engine having a diameter of cylinder 
2) = 17.7", stroke 8 = 23.6", w=160r.p.m., coefficient of regulation <J=^. The tangential effort 
diagram shows an excess area equivalent to 54 200 ft.-lbs. of work. 

Then from eq. (9), p. 218, the necessary weight of rim will be 


2900 A _ 2900X54 200 
3(Rn) 2 ~~ 1 fa(Rxmy 


Making R- 2.5 S -2.5X23.6"=4.92 ft., which corresponds to V=82.6 ft. per sec., we have 


2900X54200 
^(4.92X160) 2 


= 10 200 lbs. 


Assuming that one tenth of this weight is due to the arms, only A X 10200=9180 lbs. needs to 
be furnished in the rim. 

The cross-sectional area of rim required for this weight is from eq. (27), p. 233, 


/- 


G 


9180 


19.6322 19.63X4.92 


=95.5 sq.in. 








GENERAL ENGINE PARTS 


237 


This has been amply provided for in Fig. 314. After laying down the dimensions of the 
hub according to the empirical data above given, we proceed to compute the stresses in the 
arms. With the dimensions given in Figs. 309-316, we find these as follows: 

Tmax = .5L = .5X280 D 2 = .5X280X (17.7V 2 = 44000 lbs. 



Maximum bending moment at the hub cross-section of arm therefore is, fiom eq. (30) 

44 000X.985. 


Mb = 


6X4.92 


X 49.5 = 72 800 in.-lbs., 


from which 


10 Mb 

ab= -^r 


10X72 800 
(9.42PX6.1 


= 1340 lbs. per sq.in., 


which may be considered safe. 

The centrifugal force of one-sixth of the rim is 

„ .00034 GRn 2 .00034X9180 X4.92 X (160) 2 _ AC , nnn 1Ko 

v — 

r 


6 




































































































238 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


The tensile stress induced in the cross-section of the arms near the rim by this force is 


65000 


a = 


.7854X8.47X5.7 


= 1750 lbs. per sq.in., 


which, in view of the extreme assumptions made, may also be considered safe. 

To obtain the stresses in the joints we proceed as follows: The weight G a of each arm is 
about 550 lbs., and r 0 = 2.46 ft. Hence, according to eqs. (34) and (35a), 

C* = .00011 Rn 2 G = .00011X 4.92 X160 2 X 9180 = 127 000 lbs. , 

C a = .00068 G a r 0 n 2 = .00068 X 550 X 2.46X160 2 = 23 600 lbs.; 

2(C) = 127 000 + 23 600 = 150 600 lbs. 

Assuming the most unfavorable condition, i.e., neglecting the shrink rings and the rim 
connection altogether, each of the four 2\'' hub bolts will have to carry —-^^ = 37 700 lbs. 


This causes a tensile stress over the cross-section at the bottom of the thread of 


37 700 
2.93 


12 900 lbs. per sq.in. The real stress is of course considerably less than this. 

The two rim connections, with the same unfavorable assumption, are each loaded with 
150 600 

——— = 75 300 lbs. 

Key. This is under a bending moment 


hence the bending stress 


Jl/fc=-^-^^X3.34 = 62 goo in.-lbs.; 
4 


62 900X6 

s - per sq - in - 


There is also a shearing stress equal to 


75 300 

r= 2X1.77 X 5.55 = 385 ° Ibs ' per Sq ' in - 


K b 


With a stress ratio of « 0 = - - ~1.0 (for soft steel), the resultant stress will then be, in the 

1*0 l\-s 

most unfavorable case, 


a r = .35X7000 + .65v^OOO 2 + 4(3850) 2 = 9200 lbs. per sq.in. 


Link. This is subject to a bending moment 


73 500 

Mb =—-—X 1.85 = 17000 lbs. per sq.in., 


8 











GENERAL ENGINE PARTS 


239 


from which bending stress 


<76 = 


17000 X6 
3.15 X (3.54) 2 


= 2000 lbs. per sq.in. 


The shearing stress also occurring is 


75300 

2X3.15X3.54 


= 3390 lbs. per sq.in. 


Both of these stresses are so low that it is not w r orth while to combine them. 
Ends of the Rim. The bending moment is 


M b = 


37 650 
8 


X 1.85 =8700 lbs. per sq.in. 


and 


<76 = 


8700X0 


8.77-3.3 


X(4.91) ! 


= 790 lbs. per sq.in. 


The shearing force, considering 

, 8.77"-3.3". 

each 4.91 X--- is 


it as single shear distributed in this case over two surfaces, 


75300 oonrv 1U 

r=---- = 2800 lbs. per sq.m. 

2X4.91 ( 8 -^) 


For the case in hand (cast iron, rectangular cross-section) 


a 0 is approximately ^ = 1.3, hence 
1.0 


the resultant stress is 


o r — . 35X790 + .65^790 2 + 4(l .3X2800) 2 = 5000 lbs. per sq.in., 


which, in view of the very unfavorable assumptions made, is perhaps just safe. The pressure 
between rim and key is 

37650 u 

k= ^ — = /S00 lbs. per sq.m. 

1.77X2.73 * 1 

That between link and key is 


75300 

1.77X3.15 


= 13500 lbs. per sq.in. 













240 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


X. Governors 

1. Methods of Governing. Neglecting certain methods of regulating the speed of 
automobile engines, consisting in varying the point of ignition, entire' suppression of 
the spark, varying the size of the compression space, etc., the methods of governing 
4-cycle engines may be divided into three fundamentally different groups: 

(а) Hit-and-miss Governing (ratio of air to gas and quantity of charge per cycle, 
remain unchanged). This method may be carried out in any of the following ways: 

Keeping the fuel valve closed (operation of inlet and exhaust valves normal). 
Holding the exhaust valve open (automatic inlet valve remains closed). 

Keeping the inlet valve closed (operation of exhaust valve normal, decided 
vacuum during suction stroke). 

(б) Governing by Changing the Quality of the Charge, diagram Fig. 317 (quantity of 
charge constant and ignition regularly every fourth stroke). 

This may be carried out by: 

Changing the lift or duration of opening of the fuel valve, or changing the lift or 
duration of opening of the air inlet valve (the fuel valve being automatic); or 

Drawing back burned gases to dilute the charge (gas and air valves, or inlet 
valve, being automatic). 

(c) Governing by Changing the Quantity of Charge, diagram Fig. 318. (Fuel mixture 
of constant composition, ignition every fourth stroke.) This may be carried out by: 
Forcing a part of the combustible charge back into the suction mains; or 
Varying the moment of closing the inlet valve; or 
Throttling the charge during the entire suction stroke. 




Two-cycle engines of small capacity and also those using liquid fuels may be regulated 
according to method (a) by temporarily cutting out the gas or oil pump. For the larger 
2-cycle machines method (b) is commonly employed, adjusting the quantity of fuel gas 
furnished to the cylinder by changing the delivery of the pump (by means of throttling 
the suction, forcing back a part of the charge into the suction mains, etc.). In this 
case, if the pump does not force its charge directly into the cylinder but into an 
intermediate receiver, there is apt to be a lag of two or three revolutions before any 
given governor action commences to affect the power generated. 

The complete application of method (c) to scavenging 2-cycle engines fails on 
account of the fact that even at the minimum loads the cylinder must be completely 
filled with air, which can be only partly forced back again. Methods (6) and (c) may, 








GENERAL ENGINE PARTS 


241 


however, be used in combination in 2-cycle engines having inlet valves (not merely 
inlet ports controlled by the piston) by using (c) for the upper ranges of load, forcing 
a part of the mixture out of the cylinder, and using (6) for the lower loads, governing 
by regulating the quantity of fuel gas. 

In general, for 4-cycle engines using gas or gasoline, the hit-and-miss method of 
regulation is most economical, because it permits of the retention at all loads of the 
most favorable mixture and the most efficient maximum degree of compression. The 
method, however, possesses the disadvantage that with varying loads there is a consider¬ 
able variation in speed (see p. 228), and it should therefore be used, especially for 
the larger engines, only when the coefficient of regulation may be as high as 

In the case of kerosene and alcohol engines, hit-and-miss regulation is usually less 
efficient because, owing to continued miss strokes, the temperature of the cylinder and 
vaporizing chamber may decrease to such an extent as to lead to a condensation of 
the fuel vapor on the walls, thus causing not only a fuel loss but also fouling the 
engine. 

Methods ( b ) and (c) are usually considered “ precision ” methods of governing. In 
this respect the third method is somewhat better than the second, because in this the 
mean negative pressure (compression stroke) increases or decreases simultaneously with 
the mean positive pressure (expansion stroke), which in turn affects the variation of 
the tangential turning effort in a favorable sense. On thermal grounds ( b ) is prefer¬ 
able to (c) on account of the fact that the former at all loads works with constant 
compression, while the latter operates with decreasing compression under decreasing 
loads, and hence at smaller thermal efficiency. 

The advantage of the higher compression in the second method is, however, con¬ 
siderably overbalanced by practical disadvantages. The lean gas mixtures incident to 
the lower loads are in themselves hard to ignite, and on that account burn only 
slowly and incompletely. The heat losses due to this state of affairs are further 
emphasized when the gas content of the charge is regulated by controlling the time 
of opening of the gas valve. Owing to the shortening of the time available for 
diffusion, the naturally lean mixture is but imperfectly formed and the result is 
incomplete combustion (see Part Y). For these reasons method (c) is therefore gener¬ 
ally better than (6) also on economic grounds. Other advantages of the third method 
are that it may be carried out with much simpler mechanical means (throttle valve); 
that with decreasing compression the frictional resistances also decrease; and that, 
owing to the high vacuum in the cylinder during the suction stroke, the lubricating oil 
is more apt to reach the less accessible parts of the piston, which of course betters the 
mechanical efficiency. 

The system of speed regulation introduced by Letombe is in a certain sense a 
combination of methods ( b ) and (c). In this system the degree of compression is 
changed at the same time with the gas content of the charge, but in the opposite 
sense. Accordingly the leanest mixture is under the highest compression, which latter 
may be forced correspondingly high without danger of pre-ignition. On the other 
hand, although the richest mixture is under least compression, the ratio of expansion 
is increased, thus drawing down the otherwise unfavorably high terminal pressure and 
temperature (see description of engine in Part IV). 

Several attempts have also been made at various times to carry out the third 
method of governing with constant compression ratio,—an idea which on thermal grounds 
is quite correct. This scheme requires that the compression space must change with 


242 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


the load in order to keep the compression pressure at the same point. These attempts 
have as yet not resulted in any successful constructions, although one of the older 
automobile engines worked on this principle for some time. 1 

In order to maintain a constant thermal efficiency for all fuel ratios it is not only 
necessary to regulate the degree of compression, but the point of ignition should also 
be adjusted at the same time to correspond with the particular gas content of the 
charge. Such adjustment, when attempted at all, has so far only been made when 
disturbances in the operation of the machine showed that something was wrong, and 
in such cases only by hand. Lately, however, attempts have been made to put the 
time of ignition under governor control. It must be admitted that such automatic 
regulation of the time of ignition is very difficult to properly carry out owing to 
accidents of operation. On the contrary, hand regulation of the spark serves its 
purpose very well and is consequently coming into more and more extended use for 
the larger machines. 

2. Construction of Governors. Hit-and-miss regulation does not demand any great 
refinement in the governors used to carry it out. In them, power, stability, and range 
of movement are of almost no importance. It is necessary merely to cause the 
machine members employed to periodically interrupt the charging action to assume one 
or the other of two positions, which action usually does not call for any considerable 
expenditure of energy or constancy in the operation of the governor. For this reason 
this type of speed regulation may be carried out by simple swinging or oscillating 
members, which, having a certain freedom of movement, are forced out of their 
normal position by their own inertia. Of this type are the so-called 

Pendulum Governors. 



Of the fundamental forms of this type of governor shown in Fig. 319, form a 
shows the simple hanging pick blade, b the hanging bell-crank pick blade, and c the 
guided bell-crank pendulum. In all three forms the blade acts downward. Form d 
illustrates the bell-crank pendulum with hook latch arrangement working upward; e the 
guided bell-crank pendulum with horizontal blade; / the simple guided horizontal pick 
blade; g the double armed horizontal pick blade, and h the simple hanging pick blade 
governor working upward. Simple pick blades, just before their engagement with their 
respective valve stems, are thrown out of their normal position by interfering with 
fixed contact pieces, and are brought back into normal position at constant speed 
either by their own inertia or by spring pressure. Guided pendulums, on the other 
hand, are thrown out of their normal or engagement position by their own inertia 
only after N max has been exceeded and their deflection is consequently considerably 
less than in simple pick blade governors. 


1 See Guldner, Fahrzeugmotoren fur flussige Brennstoffe, p. 34. 







GENERAL ENGINE PARTS 


243 


Designs of Pendulum Governors: 



Fig. 320.—Gasmotor- 

en-Fabrik, Deutz. 

(At max. speed the 
weighted pendulum b, 
which is fastened to 
the valve rod a, lays 
so far behind that on 
the upstroke c fails to 
latch.) 



Fig. 321. —Pendulum Governor, Krupp-Gruson- 
werk. 

(Blade a, at Jimax, is thrown above block c 
by the action of wedge b See Figs. 238 and 
239.) 



Fig. 322. —Pendulum Govern¬ 
or, Wenzel-Zimmermann. 

(The pendulum b, carried by 
the exhaust valve rod a, pre¬ 
vents the closing of the valve 
when rimax is reached. Coun¬ 
terweight c and deflector d are 
adjustable.) 




Fig. 323.—Pendulum Governor, Crosslev Bros. 
(Pendulum b, carried by valve lever a, at Umax, lags 
behind far enough to cause blade c to miss the valve 
stem.) 


Fig. 324.—Pendulum Governor, Delamare-Deboutteville, old Type. 

(Pendulum a is carried back and forth by b. At Umax the notch at the lower end of a fails to engage 
the right hand end of the blade c. The left end of the latter drops down and misses the stem d. Fig. 237 
shows another construction due to the same designers. Here the pendulum / is moved up and down with 
the valve lever d'e. The air pressure produced in the dash pot h deflects the vertical blade to the right, and 
at rjmax causes it to miss the stem of the gas valve b altogether. Latch g holds / in this position until the 
former is unhooked in the highest position of b. The speed may be controlled by adjusting the small escape 
valve i .) 




























































































































244 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Details of Construction. To base the design of pendulum governors upon the laws 
of motion applying to the case does not offer a satisfactory solution and does not 
obviate subsequent trying out. 1 For that reason it is safer to depend upon judgment 
and feel and to provide, at least in the first models built, sufficient range of adjust¬ 
ment in those details of the construction upon which the action of the governor 
depends (length and weight of pendulum, strength of spring, angle of deflection, etc.). 
In any case, the tendency among German designers is to use the pendulum governor 
only for the smaller sizes of engine; for the larger machines they justly regard it as 
too largely subject to accidental influences in its regular operation. Delamare- 
Deboutteville perhaps went further than any other designer in the application of this 
type of governor, applying it even to his older 1000 H.P. Simplex engine. 

The fact that hit-and-miss regulation calls for but a very small expenditure of 
energy, has led to the invention of the direct acting electrical hit-and-miss governor, an 
example of which is shown in Fig. 325. In this construction a centrifugal governor, 

when the speed exceeds the normal, simply closes 
the excitation circuit of a small electro-magnet a. 
This in turn attracts the pick blade b, pulling it 
out of reach of the valve lever c, and the gas 
valve d remains closed. The governor itself thus 
expends almost no energy. 

For the “ precision ” methods of governing, only 
Centrifugal Governors need be discussed, and, in general, 
the first thing to consider in their design is the range 
of movement of the sleeve or collar, power (energy) 
of the governor being a secondary item. Numerical 
constants for power necessary, which will be gener¬ 
ally applicable, cannot be given on account of the 
variety of the constructions used; a plain throttle 

valve, for instance, usually requires less power in the governor than a movable cam, 

but more than a valve roller working on a fixed cam. It therefore becomes necessary 
either to measure the power required to move the governing mechanism and valves, 
or to estimate it. In the latter case the estimate should be at least very liberal. 

In the case of all gases not absolutely clean, especially true of suction gases, the 
governing mechanism should be so constructed that a fouling of the controlling valve 
does not interfere with the governor operation. Butterfly valves, slides, and similar 
means which give good service when used with illuminating gas, soon cause difficulties 
when used with dirty gas in that deposits of tar or other incrustations seriously 
increase friction losses and finally prevent governor operation altogether. Such occur¬ 
rences may be avoided by making the governor act not directly upon the controlling 
valve itself but upon that part of the valve gear operating this valve. In this case 
the valve exposed to the gas is positively operated by the gear and not by the 

governor. See examples of this construction in Figs. 326-329, also p. 245, etc. 

The design of the centrifugal governors proper follows the general theory. It will, 
however, hardly be necessary here to enter into any exposition of the principles 
involved, especially since the engine builder makes his own centrifugal governors only 
when compelled to do so. For, unless peculiarities of construction in the engine make 
the use of stock governors found in the market impossible, it is cheaper to buy governors 

V D ^ 6e 1894 inV “ ^ ® araz on two of the older types of Deutz pendulum governors in Z. d. 














GENERAL ENGINE PARTS 


245 


of equally good grade from builders making their manufacture a specialty. 1 In general 
it is quite easy to avoid special construction. An exception to the rule we find only 
in the smaller machines where, on account of questions of low cost and the fact that 
close regulation is not a rigid requirement, only the very simplest governing mechanisms 
aie uset - i ' ur these machines simple inertia and shaft governors (see Figs. 330-334) 
are much used and serve their purpose sufficiently well. 

Details of Some Special Governor Constructions. 



Fig. 326.— Governor Details for Large Engine, 
Masehinenbau-Ges., Nurnberg. 

(The gas valve, fastened to the wiper earn or roller 
lever a, is lifted through latch b' by the eccentric rod 
b, until the roller c breaks the connection between b' 
and a. The valve then closes under the action of the 
spring, the shock being absorbed by dash-pot f. 
Wiper cam d is supported by the bell crank e, whose 
position is changed by the governor, thus adapting 
the valve lift to suit the load.) 



Fig. 329.—Governor Construction, Gas- 
motoren-Fabrik-Deutz. 

(The governor shifts the position of the fulcrum 
about which the valve lever turns, thus suitably 
changing the lift of the gas or air valve or of the 
mixture inlet valve, as the case may be.) 




Figs. 327 and 328.—Governing Details for a Giildner 
Engine. 

(The effective lift of the inlet valve is adjusted to the load by 
shifting the fulcrum about which the inlet valve lever turns.) 



(Weight a pivoted to the larger gear of the two-to-one gearing as 
shown at n max displaces the bell-crank lever b-b' so far that the 
block b’ engages and holds the exhaust valve lever c.) 


1 Translator’s Note. This entire discussion does not generally apply to American practice, where 
there are no firms making the building of governors a specialty. 
















































































































246 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 



Weights a, in the larger gear, are connected by a link 
b. At Umax the projection c forces the lever d out far 
enough to cause the latter to engage and hold the exhaust 
valve lever. 



Figs. 333 and 334.—Shaft Governor, 6 H P. Giildner 
Two-cycle Engine. 



Figs. 335 and 336.—Suspended Centrifugal 
Governor, 12 H.P. Loutzki Engine. 


The coefficient of governor regulation o r , is often based on the difference in the 
number of revolutions per minute of the engine for the lowest, mean, and highest 
position of the governor sleeve. This agrees fundamentally with the general expression 


<\ W m ax 
Or -=— 


ttmin 

n 


^max Tlmin 

l*max T n m i n 


= 2 


2 


( ^max l^min 
^max 4" W m j n 


(1) 


Governors bought in the market usually have a value of o r varying from 2 to 4%. 
Although a small value of d r is in itself of advantage for close governing, governors 
with small coefficients are apt to act not only under changes of load but also to react 
with the variation of the turning effort within one cycle. This leads to a restless 
governor play familiarly known as “ hunting.” To avoid any possibility of its occur¬ 
rence, it is best to employ governors whose values of d r may be changed within rather 
wide limits, say from 3 to 6%. 

As a criterion of the stability of centrifugal governors the so-called centrifugal 
curve 1 has lately been extensively used, and some designers maintain that the degree 


1 Translator’s Note. By this is meant a curve showing the relation between the total centrifu¬ 
gal force of all the rotating masses involved and the distance of the center of gravity of the equiva¬ 
lent mass from the axis of rotation. A very clear exposition of the C-curves and -their use, and of 
governors in general is given by M. Tolle, in Die Regelung der Kraftsmaschinen. 























































































































































GENERAL ENGINE PARTS 


247 


of stability of a governor may be reduced to the astatic or isochronous condition when 
this C-curve is a straight line. This, however, is not correct, since in the derivation of 
this C-curve it has been assumed that the effective masses of rotation are the same 
for all positions of the governor sleeve, a condition which does not hold true for any 
of the well-known governors. On the contrary, the value for the effective mass of 
rotation changes with every different sleeve position when the influence of the arms, 
levers, etc., upon the condition of equilibrium at every different position is correctly 
taken into account. 

Hence not the C-curve itself but a “ reduced C-curve,” taking into account these varia¬ 
tions, should be used as a basis of comparison; and it is only possible to reduce the degree of 
stability of a governor to the astatic condition when the latter curve is a straight line. 1 

At any one position of the governor sleeve the total centrifugal force C must 
increase or decrease by a certain amount AC before the governor can overcome the 
resistance P at the sleeve necessary to change the position of the latter. The ratio of 
AC to C is evidently a measure of the sensitiveness or of the sluggishness of the given 
governor, as one chooses to take it. The coefficient of sensitiveness, if we may so call 
this ratio, then is 

AC (n-{-An) — (n —An) (n + An) 2 — n 2 
£ ~~C~ ~ n n 2 . 

Evidently the smaller the internal friction of the governor the smaller the coefficient 
of sensitiveness and the more sensitive the governor. In general, e should be as small 
as possible, but in any case it should be somewhat greater than the coefficient of 
fly-wheel regulation d. If this rule, which is of special importance in 4-cycle 
machines, is disregarded, the governor will hunt constantly, a state of affairs that 
can only be remedied, as a matter of last resort, by oil dash-pots or other similar 
means. The best types of centrifugal governors bought in the market have a value 
of e varying from 2.0 to .5%. If now a governor of this type is applied to an 

ordinary stock engine, in which the value of d is generally from 2.5 to 4%, the 
inevitable result is a reaction of the governor within each cycle, leading to ceaseless 
hunting, just as in the case of too small a value of d r . If there is uncertainty 

regarding the value of d, it is best to use governors which allow of some adjustment 
regarding sensitiveness. 

Concerning hit-and-miss regulation, on the other hand, a centrifugal governor can 
hardly be made too astatic or isochronous for such service because the quicker such a 
governor changes from one extreme position to the other, the smaller will be the lag 

in the action of the governor and the less frequent will be the disturbances in the 

operation of the governor during the transition periods. The limit to the sensitiveness is, 
however, in this case also set by the fact that unnecessary hunting should be avoided. 

The greatest power of regulation, that is, the ability of the governor to most 
quickly attain the state of equilibrium corresponding to a new load on the engine, is 
possessed by that governor in which the equivalent sleeve lift hr is smallest in com¬ 
parison to the actual lift. The value of the equivalent lift is 

Sum of all weights X squa re of distances moved 
Total energy 

To attain maximum power of regulation therefore it is necessary to have all masses 
to be accelerated and their range of movement as small as possible, that is, speed of 


1 See the investigations of Jahns in the governor catalogue of Wilh. Rivoir, Offenbach M. 







248 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


rotation and distance of the masses from the axis of rotation should be as large as 
possible and the path of the masses should be a straight line perpendicular to this 
axis. In the governors of usual design hr varies from ^ h (Jahn’s spring governor C) 
to h, where h is actual range of sleeve movement. 

Tables 24 and 25 are intended to give some data which may serve as an aid in 
the selection of commercial centrifugal governors for a given service. The figures given 
for the force at the sleeve possessed by the different types are based upon the 
assumption that the speed changes 2%, that is, Jn = .02n. From this the coefficient 
of sensitiveness is 

2 An 2X.02n 

e = -=-= .04. 

n n 


The energy E that must be possessed by a governor 
the sleeve may be expressed bv 

P= eE, 


so that in this case 



to furnish the given force P at 


The data furnished generally neglects the effect of governor friction upon e or E, 

that is, e is usually given too small. In more accurate terms, the available force P, 

and consequently the energy E are expressed by 

P = E(£-£r), .... (3) and E=-^— .(4) 

£ — Er 

If the shifting of the governing mechanism concerned requires an expenditure of work 
equivalent to say, A in.-lbs., the required power of a governor having a sleeve lift 
equal to h inches may then be found directly from 

hE= or E= ( A s, lbs.(5) 

In good spring governors, £ r = .005 to .02 e, although in specially bad cases the 

value may go much higher. If, in general, the efficiency of a governor is expressed by 


we may from the above also write 


P__A_ 

rj£ rjEh 


( 6 ) 

(5a) 


According to investigations carried on by Jahns and others, for a value of £ = .04, the 
efficiencies of twelve governor types examined varied between 95 and 50%, showing an 
extraordinarily wide range. 

The data given for the force P in Tables 24 and 25 are based upon the normal 
number of turns stated. If this in any given case is different from the tabular value, 
for instance, on account of operating the governor from the half-time shaft, it should 
be remembered that P, as well as A and E, vary in the ratio of the square of n. 


Example. Suppose that the work to be done in operating a given governor mechanism, includ¬ 
ing the reach rod from governor sleeve to the machine part to be moved, is found to be 56.6 in.-lbs. 
The governor is operated from a half-time shaft making 200 r.p.m. From Table 24 it will be seen 
that a Beyer governor No. 14 may be used for this work, since the power of this model is, owing 
to the reduction in speed from the normal 220 r.p.m. reduced to 

200 2 200 2 










GENERAL ENGINE PARTS 


249 


It is assumed in this ease that the governor friction is small, that is, that the 
governor efficiency ij is correspondingly high. Since it is usually possible to only 
approximately determine the resistance to be overcome, and since the resistance often 
varies considerably owing to changed conditions of operation, it is always advisable 
to choose a governor that gives considerable power over and above what seems ac¬ 
tually required. 

In certain applications of internal-combustion engines, especially where such 
engines are used to drive generators, it is often one of the requirements that the 
revolutions of the engines can be changed by hand (for instance, during the paralleling 
of alternating current generators or the charging of storage batteries, etc.). This 
requirement is generally met by the governor manufacturers furnishing some means for 
increasing the resistances encountered, as spring balances, sliding poises, etc. In such 
manipulation of the governor the effect of the changed number of turns upon the 
coefficient d of fly-wheel regulation should not be overlooked. 


Figs. 337 and 338.—Centrifugal Governor, Franz Beyer 
& Co., Erfurt. 

The models mostly used in gas engines are Nos. 10 to 
15. In each of these, both n and d may be changed 
within comparatively wide limits; the former by chang¬ 
ing the position of the nut on the spindle, the latter by 
increasing or decreasing the number of effective coils 
of the spring. For this purpose the spring plate at the 
top is so constructed that it may be turned into or out 
of the spring. The number of coils and the value of d r 
bear the following relation: 

<J r = 2 4 6 8 10% 

No. of coils required = 13.5 13 12.5 12 11.5 

This governor may also be used in a horizontal or 
in an inverted position. 



Table 24 


Beyer’s spring governor. 

. .Size No. 

10 

11 

12 

13 

14 

15 

16 

17 

Rev. per minute. 


300 

280 

260 

240 

220 

200 

180 

160 

Mean force P at sleeve. 

.lbs. 

3.75 

7.70 

13.2 

20.9 

29.8 

42.0 

59.5 

83.7 

Lift of sleeve, h . 


1.17 

1.57 

1.97 

2.36 

2.75 

3.14 

3.73 

4.53 

Power A = Ph . 


6.55 

12.10 

26.00 

49.20 

81.8 

132 

222 

380 

P 

Mean energy E ~—. 


88 

185 

338 

512 

750 

1100 

1500 

2100 

Main dimensions, inches, see 
Max. width. 

Fig. 337. 

12.25 

15.50 

19.25 

22.40 

25.70 

30.00 

34.90 

40.70 

Max. height. 

. b 

16.00 

20.30 

23.60 

26.30 

30.70 

35.40 

40.70 

47.00 

Free length of spindle. 


15.70 

17.70 

21.60 

25.50 

30.50 

35.40 

41.40 

47.30 

Dia. of spindle. 


.98 

1.18 

1.37 

1.57 

1.77 

1.97 

2.36 

2.76 

Width of groove in sleeve... 


.87 

.95 

1.02 

1.18 

1.37 

1.57 

1.77 

2.36 

Groove dia. of sleeve. 

./ 

1.57 

1.97 

2.36 

2.55 

2.95 

3.34 

3.74 

4.54 

Ext. dia. of sleeve. 

. g 

2.54 

2.95 

3.34 

3.74 

4.33 

5.12 

5.72 

6.70 

Distance from center of sleeve to lower 
face. h 

.98 

1.18 

1.18 

1.37 

1.57 

1.97 

2.36 

2.95 

Weight, without packing. . . 


44 

66 

110 

176 

253 

363 

506 

660 























































250 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 



Figs. 339 and 340.—Centrifugal Governor, Hermann Hartung 
Nachf., Diisseldorf. 

Only the smaller models, up to say No. 98, are commonly used 
for internal-combustion engines. The number of revolutions n 
of the engine can not be directly controlled, but it may be in¬ 
creased up to 15% by the interposition of a spring balance be¬ 
tween sleeve lever and governor support. This governor is also 
built with all levers completely enclosed. 


Table 25 


Hartung’s spring governor. Size, No. 

91 

92 

93 

94 

95 

96 

97 

98 

99 

100 

Rev. per minute. . 



340 

310 

240 

240 

210 

200 

190 

180 

165 

160 

Mean force P . 

.lbs. 

5.05 

7.05 

9.25 

12.50 

15.2 

19.6 

22.9 

25.3 

33.2 

50.0 

Lift of sleeve. 

.ins. 

.79 

.98 

1.18 

1.18 

1.57 

1.97 

2.36 

2.76 

3.15 

3.54 

Power.in. 

lbs 

4.0 

6.90 

10.9 

14.8 

23.8 

38.4 

54.2 

70.0 

105 

152 

Mean energy. 

.lbs. 

126 

176 

231 

312 

379 

489 

572 

638 

832 

! 1240 



’ a 

10.45 

12.20 

13.8 

15.0 

16.5 

18.1 

19.7 

21.7 

26.0 

30.8 



b 

10.90 

11.80 

14.2 

15.3 

16.5 

17.3 

19.1 

20.2 

22.3 

24.3 



c 

19.7 

21.7 

23.7 

23.7 

25.6 

27.6 

31.6 

31.6 

35.5 

39.4 

Main dimensions in inches, 


d 

1.02 

1.11 

1.38 

1.38 

1.57 

1.77 

1.77 

1.97 

2.17 

2.37 

see Fig. 339 


e 

.79 

.88 

1.02 

1.26 

1.26 

1.26 

1.34 

1.34 

1.42 

1.50 



f 

1.82 

1.89 

1.97 

2.17 

2.37 

2.57 

2.97 

3.15 

3.35 

3.55 



Q 

3.00 

3.08 

3.35 

3.75 

4.25 

4.35 

4.57 

4.73 

5.35 

5.90 



h 

1.57 

1.73 

1.89 

2.01 

2.13 

2.21 

2.25 

2.29 

2.41 

2.45 

Weight, without packing. . .. 

.lbs. 

55 

75 

115 

134 

185 

229 

286 

370 

505 

700 


XI. Ignition Apparatus 

In present day internal-combustion engines ignition is produced in one of the 
following ways: 

(a) By means of incandescent or hot tubes, and in some oil engines also through 
certain highly heated places in the walls of the vaporizer. 

(b) By means of electric spark. 

(c) By heat of compression in conjunction with the heat returned by the cylinder 
walls. This method does not require any special ignition apparatus, and does not 
therefore have to be considered here. 

1. Hot-tube Igniters. Hot-tube igniters are usually made from small tubes having 
an internal diameter varying from ^ to | in. and a length of from 2 to 3£ ins. These 
hot tubes, which may be made of porcelain, platinum, or wrought iron, are to-day in 
use only on small and medium sized gas and oil engines, but on account of their low 
cost, simplicity, and reliability are superior to electric ignition for that service. For 
gasoline engines the open flame usually employed to heat the tube may under certain 
circumstances be dangerous on account of the nature of the fuel used to operate the 
engine, while in large gas engines the hot zone of the tube is usually too far from the 
center of the charge and the hot surface is too small in comparison with the volume 
of gas to be ignited to give satisfactory service. Hence in either case electric ignition 
is better. 




































GENERAL ENGINE PARTS 


251 


Designs of Hoc-tube Igniters: 




Chimney and Burner, Loutzki Engine. 

(Chimney a together with lining c may 
be turned on the support b. Auxiliary 
air is supplied through d. e is an ordi¬ 
nary Bunsen burner.) 


one of the older Kort- 
ing Engines. 



Figs. 346-348.—Burner with Chim¬ 
ney, for the Hot Tube on a Gru- 
sonwerk Engine. 

(Boss a seats in the cylinder head, 
see Figs. 130-133, b rods supporting 
chimney, c Bunsen burner, d flame 
spreader, e slide for observing ad¬ 
justment of flame. The burner may 
be adjusted in the vertical, the en¬ 
tire chimney in the horizontal 
plane.) 


The hot tube should be so connected into the combustion chamber that the tube 
or passage can, during the suction stroke, be cleared as completely as possible from any 
burned gas, that only a good rich mixture is present at the mouth of the tube at the 
end of compression, and that the flame may spread freely in all directions and by the 
shortest possible paths. It is also essential that the inner opening of the tube be 
protected against water and lubricating oil which may stop up the narrow opening and 
in the case of porcelain tubes may lead to breakage. 

As already stated, hot tubes are to-day used only for the smaller machines, and 
the usual construction is to use them without valves, that is, in constant communi¬ 
cation with the cylinder (open hot tube). The time of ignition in such construction in 
a way adjusts itself automatically, because ignition of the entire charge can only take 
place when the velocity of the mixture entering the tube is less than the velocity of 
flame propagation of the ignition flame. 1 The latter is formed when the fresh mixture 
entering the tube has compressed the burned gases remaining in the tube into the 
'outer end and up to the hot zone. In the dead center-position of the piston the 


1 See Z. d. V. D. I., 1893, pp. 1425, 1615. 
















































































252 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


velocity of gas entering the tube has of course reached its minimum, the compression 
in the tube is the maximum, while the mixture has reached furthest into the tube. 
Consequently ignition of the charge will always take place near the dead-center 
piston position, provided of course that diameter and length of tube, position of the hot 
zone and the gas content of the charge have been properly adjusted. 

Hot-tube igniters with timing valve have almost entirely gone out of use. They 
do not offer, as shown, any marked advantage, increase the cost of the igniter gear, 
and on account of the fact that the timing valve, which is exposed to the hot ignition 
flame, is easily thrown out of order, their use simply impairs the reliability of opera¬ 
tion of the machine. It is also possible in the open hot tube to adjust the time of 
ignition within comparatively wide limits. To avoid pre-ignition, for instance, it is 

merely necessary to shift the hot zone a little toward the outer closed end of the 

tube, or to slightly contract the inner opening . On the other hand, the timing valve 
is of service in that it prevents the dangerous pre-ignitions sometimes incident to 
starting. If these are also to be avoided with the simple open tube, it is always 
necessary to start with decreased compression. This precaution of course is especially 
important in engines which compress highly during normal operation. 

The flame usually used to heat the tube is simply a Bunsen burner. The latter 
should be adjustable in two directions, so that the hot zone may be produced any¬ 
where ’along the tube, and so that the hot core of the flame may be brought further 

awa)' or closer to the tube (see Figs. 346-348). When properly adjusted the burner 
should show a blue flame and the fuel consumption per hour should be from 5 to 
7 cu.ft. of illuminating gas, from .20 to .30 lbs. of gasoline, or from .45 to .60 lbs. of 
kerosene. In the case of illuminating-gas engines, the pipe line supplying the ignition 
flame should be taken off the main supply pipe ahead of the gas bag or other pressure 
regulator (see Plate VIII in Part III). 

2. Electric Igniters. What has been said with reference to the position of hot- 
tube igniters in the combustion space, applies with equal force to electric igniters. 
Although the electric spark has a higher temperature than the hot tube, the effective 
surface is much smaller, and the proper position of the igniter is hence of even greater 
importance. Wherever the shape or size of the combustion chamber tends to interfere 
with the uniformly rapid ignition of the charge, it is better to use two or even 
more igniters properly distributed. This also tends to increase the reliability of 
operation. 

Depending upon the source of current and the igniter mechanism, either a single 
spark or a series of sparks is employed. Stationary engines usually produce their own 
ignition current by means of simple electric generators and also produce the spark by 
operating suitable make-and-break mechanism. For automobile engines jump-spark 
ignition, used by Lenoir in 1865, has of late years come into extended use. Since 
the voltage furnished by the ordinary cell is altogether too low, it becomes necessary 
\o step it up by means of induction. While automatic or mechanical interrupters 
(called tremblers) are sending a series of current impulses through the primary 
winding of an induction coil (spark coil), a series of sparks will bridge the gap 
between the two electrodes of the “ spark plug ” in the cylinder, which plug is con¬ 
nected to the secondary winding of the coil. This jump-spark, while not as strong as 
the make-and-break spark, is -satisfactory for the ignition of the rich mixtures used. 
The life of an ignition battery consisting of four cells is approximately about 400 
operating hours, which is a comparatively short time. In place of these cells special 
storage batteries are now often employed for automobile-engine ignition. With a mean 


GENERAL ENGINE PARTS 


253 


voltage of from 4-5 volts in the primary circuit, jump-spark coils require from .03 to .06 
amperes for their operation, the engine making in the neighborhood of 1500 r.p.m. 
For stationary engines, jump-spark ignition is considered unreliable and uneconomical, 
and consequently has now been almost entirely displaced by magneto ignition. In the 
traction engines of Lehmbeck an indirect magneto system of ignition has lately been 
used in which the current from the Bosch magneto is sent through a spark coil, while the 
rest of the gear is like that of ordinary jump-spark ignition. This idea is not patentable. 1 

Designs of Electric-magnetic Ignition Apparatus: 




Figs. 349 and 350.—Bosch Ignition Apparatus, Stationary 
Armature. 



Fig. 351.—Igniter Gear, Maschinen- 
bau-Ges. Niirnberg. 

(Eccentric shaft b is shifted about 
its axis by adjusting lever a, causing 
the latch c to snap off the armature 
lever at the desired instant. The 
operating rod is actuated by means of 
a crank with a throw of 1", from the 
end of the lay shaft.) 



Figs. 352 and 353. —Stand¬ 
ard Bosch Ignition Plug 
for the Smaller Engines. 

(The stationary electrode 
may also be insulated by 
encasing it in porcelain, 
soapstone or enamel.) 


1 Translator’s Note. In this country a number of different jump-spark ignition systems are 
used. See any special book treating on the subject of ignition. 































































































k-/<5!5'->t‘-/<9->I*- £7 


254 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 



Figs. 354 and 355.—Ignition Apparatus with 
Adjusting Gear. Unterberg & Co., Stuttgart. 

(Bell-crank a, instead of swinging about the 
center of the armature as usual, moves up and 
down following the cam c. By means of adjust¬ 
ing screw b, the distance between a and the ver¬ 
tical line through the armature center may be 
changed, thus varying the point of ignition.) 



Water 






Hole 




1 

\ 

□I 

" I s 

—naji? 


1 1 

_L_ 



Figs. 356-360. — Water- 
cooled Ignition Block for a 
Large Blast Furnace Gas 
Engine, Maschinenbau-Ges., 
N urn berg. 

(a, enameled stationary 
electrode made of steel; b, 
movable electrode made of 
Durana metal; c, hard rub¬ 
ber insulation; d, mica insu¬ 
lation; e, porcelain insula¬ 
tion.) 



Figs. 361 and 362.—Igniter 
Gear used by Maschinen- 
bau-Gesell., Nurnberg for 
Double Ignition. 

(When the rod a, through 
the bell-crank b, pulls the 
linkage cdd’ and the bell- 
cranks e and e' to the left, the 
dogs /and/' will turn g and 
g' through 25°, after which / 
and /' snap off. By means 
of spindle h the fulcrum 
spindle i of the bell-crank b 
may be shifted, thus adjust¬ 
ing the point of ignition be¬ 
tween — 20° and +60° of 
crank-angle. Each igniter, 
however, may be adjusted in¬ 
dependently by changing the 
positions of /and /' by means 
of the screws k and k'. 














































































































































































GENERAL ENGINE PARTS 


255 


For magneto ignition the various types of apparatus built by Robert Bosch in 
Stuttgart have practically become the standard. The latest model, originally built for 
high-speed engines, has also found extensive application to stationary engines. 

The Siemens I-armature a (Figs. 349, 350), which in the older constructions had to 
swing through an angle of approximately 50° for every current impulse produced, in 
the later type remains stationary between the poles of the built-up horseshoe magnet 6. 
The lines of force are deflected by a split sleeve c made of soft iron, which slips over 
the armature a. For this purpose it is necessary just before the spark is desired to 
turn the sleeve about its own axis through an angle of from 20 to 25% from its 
normal position, which in stationary engines is usually done by a cam on the lay 
shaft. In high-speed engines the sleeve is operated positively either by crank or 
eccentric. The turning of the sleeve puts the two helical springs shown in strong 
tension so that at the moment the cam releases the sleeve the latter is brought back 
to its original position with great rapidity. During this movement a strong current 
impulse is induced in the winding of the armature a. One end of this winding is 
grounded on the frame of the magneto itself, while the other is connected to the 
contact d. The latter in turn is connected with the insulated electrode e of the make- 
and-break plug, as shewn. The inner arm of the lever g is held in contact with e by 
means of a spring. A forked rod h is connected on one end to the vertical arm of 
the sleeve lever, as shown, while the fork rests on the pin of the outside bell crank g. 
At the instant the cam on the lay shaft releases the sleeve lever, the inner end of the 
fork strikes the pin on which it slides, and the inside arm / is suddenly brought out 
of contact wdth the pin e, which causes a spark to form between the two. On account 
of inertia effects, the time of sparking lags a little behind the instant of snapping off 
of the sleeve lever; to cut this lag down to the lowest possible amount all moving 
parts should be made as light as permissible. 

The length of the make-and-break spark depends upon the rapidity with which 
contact is broken; for that reason the inside lever should be made as long as possible, 
a good length is twice that of the outside arm g. If it is not possible to use this 
proportion, make the vertical sleeve lever (b in Fig. 364) correspondingly longer. 

The interruption of the current should be made to take place at the instant that 
the sleeve is most strongly affected by the magnetism of the poles, that is, where the 
sleeve offers the greatest resistance to turning. (This position may easily be found by 
revolving the sleeve by hand.) The fork in its normal position should not bind the pin 
upon which it slides, to make sure that / is always in contact with e. The striking of 
the pin by the fork should only be done by the sleeve lever when the latter on its 
return to the point of rest momentarily over-travels this position. The greater the 
frequency of the spark required in a given time, the smaller should the deflection of 
the sleeve be made; to counteract this the electromagnetic power of the apparatus 
must of course be correspondingly increased, a condition which should not be forgotten 
while ordering. If for some reason or other it should at any time become necessary 
to remove the armature, a piece of iron of approximately the same size should be 
put in its place to keep the magnetic circuit closed. 

The stationary electrode e requires high grade insulation (covering with asbestos, 
porcelain, or soapstone), because the ignition current possesses high voltage. Care 
should be taken to see that no short circuits can be formed through incrustations of 
burned oil, soot, etc. The spindle of the movable electrode is often made with a 
conical seat at the inner end, which serves as a packing surface (see Fig. 352). Where 
the compression is high, however, a flat seat is better, especially if the collar rests 


256 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

against a steel sleeve, because this construction minimizes the danger of the very 
troublesome sticking fast of the electrode, and the grinding in of the conical seat 
(see Fig. 360). Thorough cooling of the igniter block is of importance. The stationary 
electrode and the inside hammer lever should always be made quite large, since these 
parts if too light are apt to become red hot, which under high compression may lead 
to pre-ignition. 

Since the ignition or explosion of the charge should be completed as far as possible 
at the dead-center position, it becomes necessary to cause the spark to occur before 
this position is reached; and, in general, the leaner the mixture and the higher the 
piston speed, the more should the spark be advanced. At starting the spark should 
be so far retarded that dangerous pre-ignitions cannot take place. While this spark 
adjustment may in general be easily carried out by very simple means, large gas 
engines should be equipped with apparatus which admits of fine adjustment in order 
to be able to control the combustion with accuracy at any time during operation as 
well as during testing out. In the Unterberg magneto (Figs. 354, and 355) such spark 

control apparatus is directly combined 
with the ignition apparatus, which for 
the smaller gas engines is probably a 
desirable construction. In large gas 
engines the spark adjustment had better 
be made by changing the position of 
some parts of the ignition gear. Examples 
of this are shown on pages 253 and 
254. Automatic adjusting of the spark 
through the engine governor is beset with 
great practical difficulties, and for that 
reason has so far not found successful 
application. 

The principal dimensions in inches of 
the new Bosch magneto with exposed 
springs, for r.p.m. £175, may be taken 
from Figs. 363-365 and Table 26. 




Figs. 363-365. —Electro¬ 
magnetic Ignition Ap¬ 
paratus, built by Rob¬ 
ert Bosch, Stuttgart. 

(a, connecting post; b, 
oscillating lever; c, fork 
to make-and-break me¬ 
chanism.) 


Table 26 

Dimensions in Inches. 


Type. 

A 

B 

C 

D 

E 

F 

G 

H 

I 

K 

L 

M 

N 

O 

P 

C . 

7.10 

5.67 

3.78 

.79 

3.25 

1.77 

6.02 

7.47 

3.15 

3.19 

f 

1.97 

1.97 

6.70 

1.77 

m 3 . 

8.66 

7.20 

5.28 

.79 

4.01 

1.77 

8.45 

8.45 

4.25 

4.33 

* 

3.15 

3.15 

1.97 

7.48 

2.17 

. 


8.85 

6.90 

.79 

4.85 

1.77 

8.45 

8.45 

4.25 

4.72 


or 

4.75 

1.97 

7.48 

2.17 


These magnetos are also built with a single vertical inclosed spring. The construc¬ 
tion shown with two springs overcomes the one-sided bearing pressure in the sleeve 
support and insures better operation. 

































































GENERAL ENGINE PARTS 


257 


XII. Pedestals and Foundations 


Internal-combustion engines require foundations of somewhat better grade than 
steam engines of equal power, both because the working stresses in the former are 
considerably higher and because these stresses vary with greater rapidity. Further, the 
fly-wheel weight in gas engines is usually greater and the reciprocating parts are not 
cushioned on passing the dead center as in steam engines. All of these things combine 
to make the service required of the foundation more severe. 

Small gas engines, especially vertical machines, may up to 15 H.P. be set upon 
the floor without stone foundation, provided the weight of the engine and the vibra¬ 
tions are distributed over a sufficient number of floor beams or over a sufficiently 
large area by means of sub-bases or pedestals (Fig. 366). To deaden any resonant 



Fig. 366. —Horizontal Engine with 
Pedestal. 



vibrations, the hollow spaces in the sub-base are best packed with sand, and a sheet 
of asbestos, hard felt or cork, may be placed between the pedestal and the engine-bed 
proper. It goes without saying that larger machines which are held down in this way 
require more thorough balancing of moving parts than is ordinarily practiced in 
stationary engines. 

An engine in steady operation should whenever possible be furnished with a stone 
or concrete foundation (Fig. 367). Such a foundation is usually cheaper and serves 
the purpose better than any other. The height of the foundation above the floor 
depends upon the height desired for the crank shaft, and in engines which are 
furnished with a cast-iron pedestal the foundation usually has only a course or two 
above the floor. The height of the pedestal should be taken accordingly. 

Designs of Engine Supports or Pedestals: 


Figs. 368-370.—Pedestal for a 6 
H.P. Vertical Kerosene Engine, 
B&nki, made by Ganz & Co., 
Budapest. 

Engine D = 6.4" 

S= 10" 



(See also Figs. 45-47, p. 91.) 












































































258 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 



Figs. 371-373.—Pedestal for a 6-8 
H.P. Horizontal Engine, Fr. 
Krupp-Grusonwerk, Magdeburg. 
Engine D=7" 

S= 12' v 


Concerning the construction of the foundation, the following may be noted: In 
calculating the load to be carried by the soil, the weights of both engine and founda¬ 
tion must be considered; it must always be less per unit area than the compressive 
stress on the material of the foundation, and should not exceed 


15 lbs. per sq.in. for soft clay and fine damp sand; 

30 “ “ loam, medium hard clay and dry sand carrying clay; 

45-60 “ “ hard clay and completely dry sand; 

75-90 “ “ hard-bedded sand and gravel. 


For first-class soil the mean allowable load may in general be taken at from 60-75 lbs. 
per sq.in.; in ordinary soil it is better not to exceed 45 lbs. per sq.in. 

The material for foundations for small engines should of course be squared stone, 
where such can be had at reasonable cost, otherwise brick or concrete is used. Where 
stone is used, the mortar should contain as a binding material only hydraulic lime or 

cement, especially if the foundation is exposed to ground water. Hydraulic lime is 

cheaper and serves its purpose sufficiently well when the soil is constantly damp and 
there is plenty of time for the setting of the mortar (2-3 days). In all other cases 
hydraulic cement, Portland or Rosendale, is to be preferred. The mortar should 
preferably consist of good Portland cement and fairly coarse sand free from clay. If 

required the sand should be screened and washed, which may be done by immersing 

it in shallow wooden tanks and after a time drawing off the water. The proportion 
of cement to sand may vary from 1 to 3 (normal) to 1 to 2 for foundations carrying 
very heavy loads; for a finishing mortar for that part of the foundation above the 
floor a ratio of 1 to 1 may be used. The adhesiveness or strength of the mortar 
increases with the proportion of cement. Since natural stone does not bind as well as 
brick, it is better to use a rich mortar for the former (1 to 2 or 1 to 1). 

The quantity of water used for the mortar varies somewhat, the best proportion 
seems to be from 25 to 30% of the weight of the cement used.’ Sand and cement are 
first thoroughly mixed, afterward the water is added. For low-grade foundations, or 
for the lower third, say, of any foundation, a leaner grade of cement mortar (1:4 or 
1:6) or cement lime mortar (1 of cement, 5-6 of sand, 1 of slaked lime) may be used. 
One cubic yard of rubble masonry requires from .3 to .35 cu.yd. of mortar, while 
1 cu.yd. of squared stone masonry requires from .08 to .15 cu.yd., depending upon 
the thickness of the joints. 






































GENERAL ENGINE PARTS 


259 


Where brick is used only the hard burned should be employed; the ordinary 
building brick does not possess enough strength or durability. In the case of large 
foundations, however, the latter may be used for the inner layers or core, where it is 
protected from dampness and less heavily loaded. Each cubic yard of brick foundation 
requires about 500 brick and .3 cu.yd. of mortar, the total weight being about 
3000 lbs. 

The foundation should be carried down to solid soil and below the frost line If 

the soil does not appear quite safe it should first be prepared by laying down a 

“ gridiron sub-foundation of timbers; while dangerous soil can only be taken care of 
by a sub-foundation of piles. If there is any uncertainty as to the quality of the 
soil, it is well to start the foundation with a larger surface than required and to step 
in the successive courses until the last one corresponds to the template. Instead of 
this the bottom of the hole may also be covered with a layer of small stone broken 
brick, etc., to a thickness of from 1 to 3 feet. The repeated use of thin’ cement 
mortar soon binds this layer into one solid block, which should be larger by its own 

thickness on all sides than the foundation itself. Soil may be considered safe when it 

consists of rock or stony earth, sand or gravel (when dry), dry loam and clay (when 
the strata are at least 15 ft. thick) ; it may be considered unsafe when it consists of 
loam or clay when in thin layers or wet, quicksand, very wet sand, marl, alluvial 
deposits, etc.; and it is dangerous when it consists of turf or peat soil, swamp, filled 
.soil of any kind, etc. 

The “ gridiron ” sub-foundation above mentioned is made by placing timbers 
10" or 12" square, parallel to each other and about 4-8" apart, level on the bottom 
of the pit. The intervening spaces and the clearance at the ends are then solidly filled 
up by tamping home concrete made up of 1 part by vol. of cement, 1 of lime, 6 of sand 
and broken stone. The timbers are then cross-connected by a layer of boards 3-4" 
thick, spiked home. In pile foundations a similar grid is placed on piles which have 
been driven into the bottom. These piles are from 8" to 16" in diameter, and from 
10' to 20' long. The distance between piles and rows of piles is from 25" to 50". 
The individual piles are connected by strong cross-beams and upon these are placed 
planks as above. The position of these should be at least from 12" to 20" below 
ground water. In both cases the foundation proper rests upon the planks. The best 
timber for the piles, beams, and planks is oak, pine, or hemlock, each of which may 
first be impregnated if desired. In every case the factory furnishes the foundation 
plans. When desired the buyer should also be given directions for setting and lining 
up, which should make mention of the following points: 

After the engine is placed on the foundation and the foundation bolts are entered 
into the holes in the frame, the crank-shaft and the center line through the cylinder 
should next be accurately leveled up (paying due regard of course to any shafting 
that may have to be operated). The best way to do this is to use thin iron wedges 
which are driven under the engine frame. Care should be taken to see that the 
foundation bolts are long enough to furnish sufficient thread for nut and washer. 
After the engine is leveled the nuts are then lightly drawn up by hand. 

The next operation is to surround the entire foundation with a collar by using 
thin boards, or by piling sand or loam along the edges, and to pour in very thin 
cement mortar until the entire foundation is covered up to the level of the engine 
frame. Care should be taken to see that the mortar covers every part uniformly, for 
which purpose a piece of thin band iron may be drawn back and forth under the 
frame a few times. 



260 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

The mortar should next be allowed to set or harden completely, after which the 
foundation bolts may be fully drawn up. Care should be taken, however, to see that 
the operation does not warp the frame. For this purpose disconnect the connecting-rod 
from the crank, and keep turning the shaft in its bearings to see that it does not 
bind during the operation of drawing up the bolts. After this is completed the 
fly-wheel may be put on the shaft. ' ■ 

To facilitate the putting on of the wheel, construct an inclined plane of such a 
height that the bore of the wheel comes fair with the shaft, after which the wheel 
may be shoved into place. 

See that the bore of the wheel, the wheel seat, and the key seat are free from 
dirt or paint and oil them thoroughly. In order to prevent the shaft from turning 
while the wheel is forced into place the former may be locked by placing a block of 
wood under the crank in the crank case. 

The dimensions of foundations not only depend upon the type and size of engine, 
but also, and very materially, upon the kind of soil encountered. If the latter is of 
medium quality, still considered safe, an engine having a cylinder diameter of D 
inches, whether of the horizontal or vertical type, requires a depth of foundation below 
the floor of from 4 to 5 D. In all cases, however, the foundation should reach below 
the frost line (3 to 5 ft.). 

The length and width of the foundation of course depend directly upon the size 
of frame or pedestal. The volume of the foundations for the different types of engines, 
may be taken on the average as follows: 

For horizontal engines, without outboard bearing, 14-18 N n cu.ft.; 

“ “ “ with “ “ 21-25 N n cu.ft.; 

“ vertical “ without “ “ 7.7-8.8 N n cu.ft.; 

“ “ “ with “ “ 9.8-10.5 N n cu.ft. 

where N n = normal brake horse-power. 

It is a matter of experience that horizontal vibrations are more severe on foun¬ 
dations and soil than vertical ones, hence for uncertain or unsafe soils, vertical engines 
are better suited than horizontal machines. See p. 316, covering costs of foundations. 

Instead of having the foundation or anchor bolts pass down through the entire 
foundation, it is advisable for large engines to make these bolts shorter and thus 
to make them less yielding. For this purpose the lower end of the bolt is held by a 
cast-iron or steel plate which has been let into a recess in the foundation somewhere 
near the top. For standard sizes of anchor bolts, see page 90. 



SPECIAL PARTS FOR GAS AND OIL ENGINES 


261 


D. SPECIAL PARTS FOR GAS AND OIL ENGINES 

I. Gas Engines 

The special parts of gas engines, which include mainly mixing arrangements 
and ignition gear, are intimately connected with cylinder head and valve construc¬ 
tion, and have already been discussed, see pages 194 and 250. The auxiliary 
apparatus of gas-engine installations, such as gas producers and washers, 1 gas bags, 
gas meters, etc., will be taken up on pages 268 and following. 


II. Oil Engines 

The special apparatus to be considered under this heading also consists mainly 
of devices used to transform the fuel into a gas or vapor, to thoroughly mix the 
latter with air, and to ignite the mixture. 

1. Carburetors. We distinguish vaporizing carburetors, in which the finely divided 
air is drawn through the volatile liquid fuel on the suction stroke of the engine; 

and spraying or atomizing carburetors in which the fuel is sprayed into the passing 

air in a fine stream. The latter method of forming the mixture may also be carried 

out without special apparatus in the inlet valve or the cylinder itself. 

Vaporizing carburetors are to-day largely used for stationary machines, although 

they are large and heavy, increase the suction losses of the engine and are not 
capable of maintaining a constant mixture. 2 Since gasoline consists of a series of 
hydro-carbons, of various vaporizing qualities, it will be found that the mixture is 
rich in highly volatile vapors when the carburetor is newly filled, but the proportion 
of such vapor present decreases steadily as the level sinks, until at last there remain 
only hydro-carbons difficult to vaporize, and the result is a correspondingly lean 
mixture. The construction shown in Figs. 377—379 is supposed to avoid this difficulty. 

Spraying carburetors are not used in stationary installations to the extent that 
their low cost, simplicity of construction, flexibility and ability to furnish a constant 
mixture would warrant. 3 For automobile engines, on the other hand, their use is 
universal and they have shown themselves reliable for speeds up to 2000 r.p.m. The 
construction of these carburetors has become a special manufacturing branch of the 
general field. 4 


1 The building of power gas producers is as yet largely a branch of the gas-engine industry, 
although there are manufacturers making the former a specialty. This separation of the two fields 
is justified when the building of producers is carried on by men acquainted with the general gas- 
engine field, but leads to unsatisfactory results when it is taken up by boiler works or similar shops 
merely as a side line. One of the undesirable features certainly consists in the increased difficulty of 
meeting all of the special operating conditions of the various types of engines. This quite often leads 
to trouble at the place of erection, costly both in time and money and in many cases causes the 
engine builder to abandon the use of producers of make other than his own for some time to come. 

2 Translator’s Note. The first part of this statement hardly applies to American practice. 

3 §00 J^Q^0 J 

4 See the pamphlet of Perisse, Les Carburateurs (Paris, Masson et Cie.), which gives an extensive 
treatment of the construction and tests of carburetors, especially those used in automobile practice. 



Designs of Carburetors 



Figs.374 and 375. —Carburetor, Gasmotoren-Fabrik 
Deutz. 



(During the suction stroke the vacuum in a -null cause air 
to enter through muffler b and screen c. Saturated with 
gasoline vapor the air next passes through d and the check- 
valve e, both inserted for safety, on its way to the engine. 
Here it is mixed with more air thus forming the normal 
mixture. To facilitate the formation of vapor in o, the warm 
cooling water from the engine is circulated through'/. To the 
same end a part of the exhaust gases may be passed through 
the hollow bottom g, the adjustment being made at h. In 
winter it becomes necessary to first fill / with hot water, if 
the carbureter is located in a cold room.) 


Fig. 376.—Spraying Carburetor, Martini & Co. 

(The gasoline supply through A is regulated by means of 
the needle valve B. Air enters through C, and on the suction 
stroke of the engine the disk valve F moves downward. 
This motion causes the adjustable needle passing through F 
to free the lower end of the gasoline duct. The liquid then 
flows out through fine radial openings and is picked up and 
carried along by the stream of air. By adjusting C a part 
of the air may be allowed to pass directly into the passage 



Figs. 377-379. — Carbure¬ 
tor, Chr. Reichman, 
Munich. 

(This carburetor is designed 
to furnish a mixture of more 
uniform composition than that 
which can be obtained in car¬ 
buretors of design shown in 
Figs. 374 and 375, in which 
the more volatile parts of the 
liquid always distill off first. 
In this design the gasoline is 
fed drop by drop from the 
reservoir a through the cen¬ 
tral pipe and is absorbed by 
the suspended wicks 6. From 
these it evaporates and mixes 
with the passing air. The 
wick chamber is surrounded 
by a water jacket c, which 
contains a nest of tubes d 
through which the exhaust 
gases coming through muffler 
e may be passed. Chamber/ 
is filled with pebbles or with 
wire screens and acts as a 
safety device.) 


262 












































































































































































































































































































SPECIAL PARTS FOR GAS AND OIL ENGINES 


263 


Fig. 380.—Spraying Car¬ 
buretor, Sintz. 

(Valve a operates 
needle valve b. By ad¬ 
justing c, a part of the 
air may be led directly 
into the suction pipe.) 




Fig. 381.—Double Carburetor, Banki. 

(The air drawn in by the engine passes through 
a pipe located between the two float chambers. 
Gasoline and water are supplied to this pipe by 
means of from two to four spray nozzles.) 


2. Vaporizers and Atomizers. Both of these are intended to change the heavier 
liquid fuels, as kerosene and alcohol, to a state in which they will readily mix with 
air. In the former the oil is vaporized with the agency of heat, and the oil vapor 
is either immediately or subsequently mixed with air. Generally the oil is previously 
sprayed or atomized in order to spread it over a large surface. The well-known fact 
that liquids vaporize more easily in a partial vacuum has to the writer’s knowledge 
never been practically utilized in oil engines. 

If the oil is so completely atomized as to allow of the suspension of the particles 
in air, which, with suitable means may be easily accomplished, it is quite possible to 
form a satisfactory and uniform air-oil vapor mixture without the aid of heat. The 
oil thus mechanically carried along vaporizes only after reaching the cylinder during 
suction and compression. In order to get a clear insight into the action of vaporizers 
and atomizers, as well as oil engines themselves, it will be well to briefly consider the 
vaporization phenomena. 

As shown in Part IV in greater detail, the crude oils are composed of a series 
of hydro-carbons having very different boiling points. The most volatile of them 
readily vaporize under temperatures ranging from 60° to 70° F., the greater propor¬ 
tion, however, require much higher temperatures for vaporization. Taking refined 
kerosene, for instance, usually more than 60% by volume will vaporize between 400° 
and 570° F., while solar oil loses 75% between 470° and 630° F. (see Fuels in Part 
IV). The degree of heat required to vaporize the heaviest constituents of the fuel oils 
is not far from a dark-red heat. We may assume that the temperature of the cylinder 
walls of an oil engine is somewhere between 140° and 340° F., the average probably 
in no case exceeds 300° F., but that is considerably below the condensation tempera¬ 
ture of most of the constituents of the fuel oils. If therefore oil vapor, which has 
been produced in a special chamber or vessel sufficiently heated, is introduced into the 
engine cylinder, it is unavoidable that a part of the hydro-carbons of high boiling 
point in the vapor will, owing to its contact with the cold air simultaneously drawn 
in and its contact with the cylinder walls, returns to the liquid state and will be 
thrown down in the shape of drops. The reason for this occurrence of course lies 
in the fact that both the quantity and the specific heat of the oil vapor are too small 
to maintain the required mean vaporization temperature against the cooling effect of 
both air and walls, and a partial condensation of the oil vapor is consequently inevita¬ 
ble. The liquid thrown down burns only partially, owing to the fact that air can 
only reach its exterior surface, it gives off vapors during the exhaust stroke and 
finally forms incrustations on the inner walls. 

“Superheating” the oil vapor does not have the desired result, first because the 



























264 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

degree of superheating is limited on account of low temperature of ignition of the 
oil, and again because whatever degree can be employed has little effect upon the 
mean temperature for reasons already stated, and hence cannot in any marked degree 
prevent condensation. The scheme of superheating the oil vapor, patterned without 
good reason after steam-engine practice, has therefore nothing in its favor. 

On the other hand, the opposite of this scheme that is to introduce the oil 
mechanically very finely divided (in shape of “fog”) into the air and thus into the 
cylinder, promises better success. The oil “fog” thus produced by means of com¬ 
pressed ’ air or other medium is very similar to the ordinary oil vapor with the 
difference, however, that it possesses ordinary room temperature and is much more 
stable than the latter (a cloud of oil particles so formed will remain suspended in 
air for several seconds and only gradually disappear.) Condensation is entirely 
avoided since the oil in effect retains the liquid state. It is of course im¬ 
possible to prevent some of the oil particles from gathering into drops when 
striking valve disks and cylinder walls; but intelligent design of the inlet passages 
can reduce this difficulty to a minimum, a thing which is very difficult to do with 

oil vapor. , * 

Success should therefore be sought in complete atomizing and spraying rather 
than in high temperature vaporization. This conclusion is further justified when we 
consider the subsequent steps in the combustion process. Heated oil vapors soon 
reach their ignition temperature during compression. This means the use of a lower 
compression pressure, which, in turn, as has previously been pointed out, leads to lower 
efficiency and smaller specific capacity of engine. The latter loss is aggravated by 
the fact that the density of the heated charge is low at the outset. On the other 
hand, the transformation of the oil “fog” into oil vapor during the compression stroke 
utilizes some of the heat of compression, hence a higher, that is more efficient degree 
of compression becomes possible. This advantage is very strongly apparent in the 
case of alcohol vapor with which, on account of the high specific heat of the water 
carried and the consequent lowering of the temperature, compression pressures of from 
10-12 atmospheres may be carried without pre-ignition. The general axiom “to have 
the mixture as cool as possible ” therefore applies also with equal force to oil 
engines. 

The vaporizer is placed either before or behind the inlet valve. In the latter case it is 
apt to lose its separate identity. It is heated either externally, by a special heating 
lamp, or internally by the heat of combustion. As long as the vaporizer temperature 
remains considerably below the ignition temperature of the mixture, the oil may be 
vaporized with unlimited air supply, i.e., the mixture may be formed at the same 
time. Highly heated vaporizers on the other hand should contain little or no air, 
as the mixture may take fire on the hot walls, consequently the mixture can only 
be formed beyond the vaporizer proper. It is hardly possible to make a statement 
of general applicability regarding the superiority of one or the other of these two 
types of vaporizers. The writer believes, however, that the moderately heated vapor¬ 
izers will come into use largely for engines using kerosene only, because a colder and 
more uniform mixture may be obtained from them than from the highly heated 
vaporizers. For alcohol engines, low vaporizer temperatures always serve the 
purpose, since alcohol starts to vaporize at about 175° F. and water at 212°. 
The widespread impression that alcohol requires an especially high vaporizer 
temperature is therefore without foundation. 


SPECIAL PARTS FOR GAS AND OIL ENGINES 


265 


Designs of Vaporizers, Atomizers and Heating Lamps: 



Oil 



Air 



Fig. 382.—Capitaine Vaporizer. 

(The liquid fuel enters the tube through a very 
small opening. The tube itself ends in the com¬ 
bustion chamber, just below the inlet valve.) 



Fig. 383.—Vaporizer, Henriod Schweizer 
(Lude Type). 

(The inlet valve a has two disks, of which b con¬ 
trols the mixture and c the quantity of oil. By 
means of screw d and sleeve e the capacity of the 
space above c may be regulated.) 



Exhaust Gases-. 

Fig. 384.—Vaporizer, Ivorting Bros. 



Figs. 385 and 386.-—Durr Vaporizer. 



Figs. 387-389.—Kjelsberg Vaporizer, Ludwig Nobel, 
Petersburg. 

(Inlet valve a operates the fuel valve c through the link¬ 
age b. The kerosene admitted by c drops from the atom¬ 
izer plate d on to the radial webs of the vaporizer e, through 
whose jacket space e' the hot gases from the hot tube heat¬ 
ing lamp / are passed, f is the burner, h a small basin for 
pre-heating/ , , i a screw to adjust the hot zone along the 
tube g, k the needle supply valve for the burner.) 







































































































































































266 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 



Figs. 390-393 —Priestman Vaporizer. 


(The atomizer (b in Fig. 395) sprays into the annular chamber 
a, while the air supply pipe (c in Fig. 395) opens into the vapor¬ 
izer chamber b. Jacket space c is heated by the exhaust gases, 
while the passages d are heated by a heating lamp at starting.) 



(The governor, by regulating the position of the plug-cock a, 
controls both the supply of oil through b and of air through c. 



Fig. 396. —Petreano’s Carburetor and 
Pre-heater. 

(Air and fuel oil are supplied to the 
upper space a. The oil is absorbed by 
metallic wicks b, which cover the central 
tube, is vaporized and picked up by the 
passing air. The mixture passes through 
chamber c to the engine.) 



Fig. 397.—Kerosene Heating 
Lamp, Capitaine. 




Figs. 398 and 399.—Hornsby Atomizer Plug and Governor Valve. 


(The action of the atomizing plug or valve a is automatic. The by-pass 
valve b is opened more or less by the governor, through c, so that a variable 
part of the oil furnished by the pump flows back into the reservoir through 
d.) 


r : ] 

Atomizer. 

Plate 


Cooling Water 


Petroleum 


3. Liquid Fuel Pumps. The minuteness of the quantity of oil to be furnished to 
each charge requires pumps of very special construction, some types of which are 
shown in Figs. 400-405. 












































































































































SPECIAL PARTS FOR GAS AND OIL ENGINES 


267 


Designs of Fuel Pumps: 



Pump. 


(The plunger is operated by the 
inlet valve lever.) 



the suction and discharge 
passages are opened and 
closed by the cylinder a slid¬ 
ing up and down in the hous¬ 
ing b. 



& Co 



Fig .403. —Oil Pump, in which 
the inlet and discharge 
parts are controlled by the 
small piston or plunger a. 



Fig. 405. —Oil Pump and Governor Details, Small Augsburg- 
Diesel Engine. 

(Pump plunger a is operated from the lay shaft h. The latter, 
through the guide g, also operates the lever e, to which the link¬ 
age d of the suction valve c is fastened. Discharge valve b is 
automatic. As the governor varies its position, depending upon 
the load, the pivot / of the lever e changes so as to cause the 
suction valve c to close at varying points in the discharge stroke 
of the plunger a, thus controlling the amount of oil delivered to 
the cylinder.) 

























































































































268 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


In order to avoid too minute dimensions for the piston, the displacement of the 
pump cylinder is often made from 3 to 6 times that required by the maximum 
engine capacity. The proper proportion of the mixture is then obtained by allowing 
the plunger to work idle a part of the stroke, forcing a part of the oil through the 
suction valve or a special by-pass into the supply reservoir. By changing the instant 
at which this return is interrupted the quantity of oil supplied the machine can be 
adjusted to suit the load on the engine and thus to regulate the speed. In small 
engines the tendency at present is to do away with the use of an oil pump altogether 
and to draw in the oil by means of the engine piston itself. In such installations 
the oil reservoir is placed 1.5 to 3 ft. above the point of exit from the supply pipe 
into the cylinder, unless the oil tank is under air pressure. In order to prevent too 
great variations in the composition of the mixture when this scheme is used, the 
hydrostatic pressure on the orifice is often maintained constant by float or overflow 
valves; that is, it is rendered independent of the level of the supply in the tank. 
Good fuel pumps, however, insure greater reliability of operation, better atomizing of 
the oil and the formation of a more uniform mixture. They should therefore always 
be used in large engines. 


E. AUXILIARIES 

I. Power Gas Installations 

For information regarding fuels and power gases, see Part IV. Concerning the 
theory of the production of power gas, see Appendix. 

The general arrangement of the main parts of a gas-producer installation, that 
is, of gas producer and purifier or washer, depends primarily upon the properties of 
the fuels to be gasified. The main points to consider in this regard are the moisture 
content, proportion of earthy and tar-forming constituents, tendency to coking and 
clinkering, size of the fuel and the heating value of unit volume. Unless there is 
some counteracting agency, a high moisture content of the fuel reduces the tempera¬ 
ture of the producer in too great a degree, and in this way interferes with the gasi¬ 
fication process. Too great a percentage of ash quickly clogs up both generator space 
and grate, and hence hinders the passage of air and the development of heat. Tar¬ 
forming gases cause serious disturbances because the tarry hydrocarbons condense in 
pipe lines and valves and form soot in the cylinder during combustion. Fuels that 
show a strong tendency to coke or clinker, and those that are too small sized, inter¬ 
fere with the operation of the producer in much the same way as excessive ash. 
Large-sized or coarse fuel does not burn with the desired uniformity, offers too little 
surface for gasification, and allows air, v r ater vapor and carbon dioxide to reach the 
upper part of the producer without decomposition. Naturally those fuels most free 
from the disadvantages enumerated above, and especially those low in tar-forming 
constituents, are commonly used for the production of power gas. The fuels that 
most nearly fulfil these requirements and are approximately free from tar are hard 
coal, or anthracite, and coke; we therefore find that most gas producer installations 
are arranged to utilize these fuels. 

The method of producing a suitable power gas from bituminous fuels, especially 
from the more recent coal formations (lignite and peat) is at present being developed 
with a great deal of energy, but until now no completely practical and thermally 
efficient producer has been found. Several installations of this type now' in 



AUXILIARIES 


269 


operation have, however, shown a promising beginning. After it had been demon¬ 
strated that it was impossible with simple means to continuously and thoroughly 

remove any considerable quantity of tarry gases from the producer gas leaving the 

generator, the solution next attempted was to so conduct the gasification of the solid 
fuel that the tar could be dissociated or broken up into fixed gases inside of the 
producer itself, and thus to prevent the tar vapor from appearing in the producer 
gas at all. Tar may be broken up into fixed gases by leading the gas carrying it, 

or the tar vapor by itself, through a layer of incandescent coke together with air 

and water vapor. This either burns the tarry hydrocarbons to carbon dioxide (CO 2 ) 
and water vapor (H 2 O) or breaks them up into carbon monoxide (CO) and hydrogen 
(H 2 ). The gasification of the solid fuel and the subsequent fixation of the tar may 
be carried on either in the same generator or two or more generators may be con¬ 
nected in series above or behind each other, as shown later by several examples. 

1. The Power Gas Producers (Generators). Power gas producers are closed vertical 
retorts in which suitable fuels are burned with insufficient air supply and with the addition 
of superheated steam. The final product of. this operation is a combustible, composite 
gas containing, besides nitrogen, mainly carbon monoxide (CO) and hydrogen (H 2 ). 
In the generator proper, about 85% of the heat of the fuel is transferred to the 
gas, that is, is recovered. Starting, stand-by periods, incomplete combustion, etc., cause 
a loss of from 10 to 15% of the fuel. Where separate boilers are used, about 8% 
of the total fuel is used for the making of steam. Consequently the real efficiency 
c-f the generator is only from 60 to 75%, and only in isolated cases higher than 
the latter figure. One pound of coal gasified usually produces from 70 to 80 cubic 
feet of producer gas with an average heating value of 135 B.T.U. per cubic foot. 

Along with the same weight of coal there is introduced about .75 lb. of water, except 
in the case of suction gas generators, where for the purpose of cooling the grate the 
quantity used is generally much greater, up to 1.5 lbs. per pound of coal. 

Depending upon the method of moving the air-steam mixture through generator 
and washer, we distinguish three types of producer installations: 

(a) Pressure gas installations, in which the air and steam arfe forced through the 

fuel column by means of special blowers, and the steam is usually generated in a 

separate boiler. 

(b) Suction gas plants in which the suction of the engine piston draws the steam 

and air mixture through the charge in the generator, and in which the required 

steam is usually produced by recovering some of the heat of the generator. 

(c) Combination plants so constructed that a fan or blower draws the gas out 
of the producer and forces it towards the engine. 

Suction gas plants have almost entirely displaced the older pressure installations 
for gas engine work. They do away with the use of a separate boiler with its 
attendant disadvantages (increased fuel consumption, greater cost of attendance, 
higher capital cost, increased floor space required, insurance, inspection, etc.). 
Further it is possible to clean the grates in suction gas plants without interrupting 
regular operation, and the entire installation is more simple as well as cheaper both 
in construction and operation. Since the rapidity of gasification depends directly upon 
the suction of the engine piston, the quantity of gas made will vary directly with 
the amount required by the engine. This automatic regulation makes the use of 
gas holders or pressure regulators superfluous, and thus eliminates one of the least 
desirable features of the pressure gas plant. 

Operation with suction gas does not in itself possess any serious disadvantages, 
where real difficulties crop up they are usually due to poor construction or wrong 


270 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


management. The partial vacuum existing in the system may at first sight appear 
a serious drawback, but examination shows that with proper construction the pressure 
can be kept to within from 4" to 8" of water of the atmospheric pressure, which is 
less than the regular variations in the barometric readings. Only with very poor 
construction or bad operating conditions does the vacuum reach a degree (40" or 
more of water) that markedly affects the specific capacity of the engine. The mis¬ 
placed apprehensions concerning the danger of the vacuum existing in suction systems 
are directly responsible for the fact that the scrubbing and washing apparatus for such 
plants are made much smaller and less complete than experience has shown is neces¬ 
sary for pressure installations. As a matter of fact, suction gas may be just as 
thoroughly cleaned as pressure gas, but, in place of simply transferring and utilizing 
old designs of working apparatus it is necessary to employ washers adapted to the 
particular purpose. With this precaution it should be possible to satisfactorily clean 
suction gas without materially affecting engine capacity. 

The limitations in the use of suction gas for engines sometimes found where the 
gas is also used for other purposes (heating furnaces, etc.) have been eliminated by 
the use of combination systems or by drawing the gas intended for heating purposes 
by means of blowers and forcing it to the furnaces. 

There remain to be considered two things which nave oeen used as charges 
against the suction gas principle, serious enough to arouse insurance companies and 
to cause government supervision in some quarters—that is, dangers due to explosions 
and to poisoning. Cool and deliberate consideration of all the facts in the case will, 
however, show that suction gas plants not only warrant greater safety against either 
one of these accidents than pressure gas producers left so long without molestation, 
but further, that a number of serious and inexcusable irregularities in attendance 
and maintenance will have to concur in order to produce conditions that will lead 
to really dangerous poisoning or explosion accidents in suction plants. 1 

At the same time it cannot be denied that there are certain further problems in 
the development of suction gas plants. The number of different kinds of fuel that 
may be properly gasified in them is at present very small, and even for these there 
are features that call for improvement. The first of these is the present method of 
producing and introducing the steam because the ordinary vaporizers all labor under 
the disadvantage of having too great a w r ater space. The large amount of water in the 
vaporizer (superheater) seriously increases the time of starting, since the water must 
be brought to the boiling point before the gas has the normal composition. As long 
as the water does not steam, the gas will lack its main heat-carrying element—hydro¬ 
gen. The engine is started with difficulty and only after a considerable length of 
time may it be fully loaded. After the engine is stopped the steam finds its way 
from the vaporizer into the room, and causes trouble through condensation, rusting, 
etc. Specially bad, however, is the fact that with vaporizers having a large water 
space the composition of the gas cannot be controlled with any degree of certainty. 

1 T° sa y a f ew words in explanation of this statement it may be noted that the gas escaping 
through small leaks in pressure gas plants, even when in considerable quantities, is not noticed by the 
attendants, because carbon monoxide gas is odorless and colorless. The danger of poisoning and 
explosion in this case may therefore appear very suddenly and without warning. In suction-gas plants 
this condition cannot occur. Small leaks may of course cause a decrease in the quantity of gas in the 
mains, w ic is in most cases quickly noted by a decrease in the engine power, but to produce 
explosive mixtures requires such serious defects in piping that the condition is hardly conceivable in 
practice, because the ratio of air to gas must be in the neighborhood of 1:1. On the basis of such 
tacts the suction-gas producer had better be left unfettered with any government regulations in its 
development—it is no more dangerous than the ordinary heating stove. 



AUXILIARIES 


271 


A change of the water supply to the vaporizer has no effect, and very often produces 
the very opposite from the result intended. If, for example, the hydrogen content 
of a lean gas is to be raised, the attendant naturally increases the supply of water 
to the vaporizer. This, however, is immediately followed by a cooling of the water 
in the vaporizer as well as the vaporizer walls. The result is a reduction in the 
quantity of steam evolved and the gas made is even poorer in hydrogen than before. 
If, on the other hand, the engine shows by pounding and other signs that the hydrogen 
content in the gas is too high, the only remedy available is to cut down the water supply 
to the generator. The inevitable result of this, however, is that the evolution of steam 
becomes stronger than it was before and will remain so until the vaporizer goes dry. 

These difficulties are avoided if there is no water storage capacity whatever in 
the vaporizer, and only as much water as can be used per suction stroke of the 
engine is furnished to the generator. In case the engine is governed by throttling, 
the periodic introduction of water should be so regulated that the resulting gas shall 
always have its best composition for all loads on the engine. This requirement, 
important alike from the standpoint of economy and reliability, is to-day fulfilled by 
only a few of our suction gas plants, while all other installations operate normally 
only in the neighborhood of certain loads and are likely to give trouble when this 
load is materially changed. 

The usual construction of tube vaporizers should also be quickly changed to 

something better. In them, short tubes are usually held between inelastic tube sheets 
and constantly exposed to the effects of widely varying temperatures. The result is 
that they are often the cause of costly repairs and troublesome interruptions in 

operation. On account of the sudden decrease in the gas velocity and the change 

in direction of flow, dust is deposited thickly in the end spaces of the vaporizer 

above and below the tube sheets, not only decreasing the efficiency of heat transfer, 
but after a time seriously hampering the flow of gas. Besides this, the tubes increase 

the distance the dirty gas has to travel and their surfaces are apt to become coated 

with deposits of dust and tar, more serious in this place because of the reduction of 
heat transfer and the difficulty of proper cleaning. In the smaller installations the 
generator vaporizer may be placed directly in the top of the producer; in larger 
plants its proper place is outside, but so close to the shell of the producer that the 
distance from the producer to the scrubber shall be as short as possible, and so 
constructed that there shall be no chance for deposits of tar or dust. 

The vaporizer should be heated only by means of the waste heat of the genera¬ 
tor or the sensible heat of the gas. To use the exhaust gases of the engine for this 

purpose is from the thermal standpoint mostly folly, 1 and in many cases even abso¬ 
lutely wrong because it increases the cost of the plant as well as the back pressure 
on the engine. The advice to use a part of the jacket water to feed the vaporizer 
is equally ill-considered because for every pound of coal used the engine requires from 
60 to 90 lbs. of jacket cooling water, while only from .7 to 1.5 lbs. of water, i.e., only 
about 1/70 of the former amount is fed to the generator for the same amount of coal. 

1 Proof. Assuming a very good efficiency for the generator, say 75%, we find for every pound 
of coal (14 300 B.T.U.) gasified in round numbers 3500 B.T.U. are discharged in the shape of radiant heat 
and sensible heat in the gas. This heat is quite sufficient to evanorate about 3 lbs. of water, that is, 
more than twice the maximum amount that any normal generator would require. Any sensible heat 
that is not removed from the gas before it enters the washing apparatus is in most cases absolutely 
lost. To recover the sensible heat of the gas by pre-heating the air for the generator is possible only 
in limited degree, even in large plants fitted with special pre-heaters. The ordinary constructions of 
this type at any rate show no added advantages both on account of the increased suction and 
exhaust friction losses and because of the rapid destruction of grates, etc. 



272 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Designs of Gas Producers. (Figs. 406-431, producers for anthracite and coke, Figs. 
432-440, producers for fuels carrying tar.) For data concerning practical operation, see 
Part III. 

{a) Pressure Gas Installations 



Fig. 406. —Pressure Gas Installation, Korting Bros., Hannover. 
(For floor dimensions for complete plants, see p. 341.) 


The poor gas made during starting escapes through valve b and purge pipe a. During this time the pro¬ 
ducer operates like a common stove. As in the case of a stove, the introduction and firing of the first fuel is 
done through the open ash pit doors. After sufficient fuel is added, these doors are closed and the steam 
blower is started. In case the natural draft available is not sufficient to properly start the fire, a small steam 
ejector in pipe a may be used to increase the draft. As soon as the flame at the try-cock shows dark red, 
not blue, the gas is of proper quality, the valve b is closed and the plant is ready for operation.—The cleaned 
gas on its way to the engine passes through a regulator (marked “gas holder”). The rising and falling of 
the bell of this regulator opens and closes the steam supply valve d by means of the chain connection c-c, 
thus controlling the amounts of air and steam furnished the producer.—According to the builders, the average 
composition, in per cent by volume, of the gas made in producers of this type is as follows: 

Hydrogen, 18%; carbon monoxide, 26%; hydrocarbons, 2%; Carbon dioxide, 7%; and nitrogen, 47%. 
The efficiency of the producer is from 80 to 82%. 



Fig. 407. Pressure Gas Producer, Gasmotoren-Fabrik Deutz. /////?./ 

tbe ° . su P er ^eated in the coil located in the upper part of the boiler. During starting 

economizer c. ^opening b. The air injected by blower a is pre-heated by passing through th! 


























































































































































AUXILIARIES 


273 


Fig. 408. —Pressure Producer Gas In¬ 
stallation without Boiler, Lencauchez. 

Water is sprayed through pipe a con¬ 
tinuously against the ribbed plate b. 

It drips from here into the cast iron 
ash-pan c, vaporizing in its course. At 
the same time the positive blower d 
blows air, slightly compressed, through 
several radial channels e into the ash 
pit c where the air picks up the vapor 
and the mixture then passes through 
the producer. A fire brick arch / di¬ 
vides the charging space g from the gas 
outlet space h. The former is filled with fresh fuel which slowly moves downward as demanded. Since the 
blower is driven by the engine and no gas can be made unless it operates, the gas holder should always con¬ 
tain sufficient gas when the engine is stopped. 



X 

~r 

□ 

\ 







(b) Suction Gas Plants 

Figs. 409-411.-—Suction Gas Plant, 25 H.P., made by Societe des Moteurs 
Gazogene Benier, Paris. 

(One of the first type of suction gas producer.) 






-1 le) ^ 








The lower end a of the producer is closed by 
a hollow cylindrical grate b which in the interior 
acts as the vaporizer chamber. The latter is 
constantly supplied with water through c, any 
excess flowing out through c'. The steam formed 
reaches the chamber e through d, and finally 
finds its way into the mixing chamber /. On ac¬ 
count of the sucking action of the charging pump 
of the engine, air also enters / from the outside, 
and the air-steam mixture formed is drawn into 
the space g, down through the annular space h 
and thus reaches i. Here it surrounds b and then 
enters the producer proper. The gas made passes 
out through two washers k and k'. 

The grate may be rotated on its axis, the con¬ 
struction is plain from the figure. Purge pipe l 
is closed by a disk valve V. Charging hopper m 
is closed at the bottom by a slide n, which may 
be shifted to one side. 




























































































































































































































































































































274 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 



The vaporizer, in which the water 
is kept at the same level by means 
of an overflow, surrounds the top 
of the producer. On the suction 
stroke of the engine air enters the 
vaporizer through the elbow at the 
right, and is saturated with water 
vapor. The air-steam mixture then 
passes through the pipe at the left, 
under the grate and through the in¬ 
candescent column of coal, in which 
the well-known reactions resulting in 
producer gas take place. The gas 
next passes through the valve 5, up 
through the wet scrubber, through 
the gas reservoir and on its way to 
the engine finally through a dry 
scrubber (wire brush). 

The double lock construction of the charging hopper is plain from the illustration. On starting the pro¬ 
ducer, valves 1 and 2 in the pipe at the left of the producer are closed, while valve 3 is opened. Next to the 
scrubber, valve 5 is closed, at the same time opening valve 4 leading to the purge pipe. The hand blower 
then forces air through the producer and up the purge pipe. During stand-by periods, valves 1, 3, and 5 are 
closed, while, on account of the draft through the purge pipe, a little air passes in at 2, keeping the fire alive. 
The check valve in the pipe below valve 1 is intended to prevent any striking back of hot gas when the 
change from normal operation to stand-by is being made. Any “overpressure” of gas that may form at this 
particular moment is taken care of by an equalizer pipe leading from the seal-box to the purge pipe. (The 
construction of the larger Deutz plants does not quite agree wuth Fig. 412.) 


Fig. 412. —Suction Gas Plant, made 
by the Gasmotoren-Fabrik, Deutz. 



Fig. 413.—30-35 H.P. Suction Gas Installation, Korting Bros. 

The tubular vaporizer, located at the side of the producer, is heated by the outgoing gases. The purge 
pipe branches off beyond the vaporizer, so that the latter is also being heated during the starting period. 
The starting fan blows into the gas space of the vaporizer. (Type of producer for bituminous fuels, see Fig. 
433.) 






















































































































I 



Figs. 414 and 415.—100 H.P. Suction Gas 
Plant, Jul. Pintsch, Berlin. 

Cover and filling hopper are supported on 
a sector-shaped plate which may be swung 
about a pivot at the side. FiG. 414 shows the 
operating position. The producer is charged 
by swinging the hopper over the center open¬ 
ing, see Fig. 419. The small branch pipe near 
the bottom (at the left) of . 
the air supply pipe serves 
to furnish air under slight waft 

pressure for starting. te# 


Pressure 

Regulator 




Fig. 416.—General View 
of a 16 H.P. Pintsch 
Suction Gas Plant. 


275 





















































































































































































































































276 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 



Fig. 417. —Suction Gas Installation, Pintseh. 

(2 Producers, 600-800 H.P.) 

The gas is led through one 20" main pipe to three twin cylinder engines, each rated at 200 B.H.P. A 
Pintseh pressure regulator is connected in each branch main to the engines. The apparatus in the back¬ 
ground is a small water-gas producer (1500 cu.ft. per hour), the gas being used for heating purposes in the 
works. 





















AUXILIARIES 


277 



Fig. 418.—Scrubbing Apparatus, 600 H.P. Pintsch Suction Gas Installation. 

The picture shows in the left foreground the dry purifier for the water gas. Directly behind this is 
located the cylindrical scrubber for suction gas between two cylindrical coolers. The latter were here 
employed because the otherwise necessary cooling water was hard to obtain. At the right are two dry 
purifiers, 17J by 21 ft. floor-space, for suction gas. 



























278 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 




(D 3 Wh 


Fig. 419.—Charging Floor of the COO H.P. Pintsch Suction Gas Installation, shown on pp. 276 and 277. 

















AUXILIARIES 


279 




Figs. 420-422.—25 H.P. Suction Gas Installation, Wiedenfeld & Co., Duisburg. 


The vaporizer is located around the top of the generator, as in the Deutz 
design. The air-steam mixture is superheated by passing through the hollow 
circular grate support a. The gas main b dips under the surface of the water 
in the seal-box c, t8 throw down the coarser impurities. Overflow pipe d 
from the scrubber discharges into this seal-box. Water is supplied to the 
vaporizer through e, any overflow is led through pipe / under the grate. 


Fig. 423.—100 H.P. Suction Gas Plant, 
Dunker & Spielter, Hannover. 

The jacket space of the producer is used 
as the vaporizer, a special superheater 
coil surrounding the top of the pro¬ 
ducer. At starting a small amount of 
water is introduced directly into the coil, 
to obtain sufficient steam from the outset. 
Only a part of the air used passes through 
the coil, the rest entering the ash pit 
directly without being mixed with steam. 







Seafi ! 

i 





S 

Box | ! 

i 


-JPcSC/ - x-0 <?—J 


































































































































































































280 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


j 



Figs. 424-427.—30 H.P. Gtildner Suction Gas Instal'ation. 

(.C uts show the plant as arranged for starting up.) For floor space required for complete plant, see p. 356. 

The producer has no vaporizer containing any quantity of water; air is pre-heated in the top space a, then 
saturated with steam through the action of the supply mechanism a' (see Fig. 428). The mixture then flows 
through superheater b, passage c and safety valve d into the ashpit e. The lining / is cone shaped at the 
bottom. The gas passes through double valve h, passages g and h into the wet scrubber i; starting fan k 
blows through pipe l into the ashpit. Overflow m of the wet scrubber discharges into the sealbox n. 



Action of the Water Supply and Regulating Valve, Fig. f28: At the center of 
the water pan a there are two concentric tubes b and c, forming an annular space 
d between them, which is in communication with a through openings b r . As 
long as the apparatus is at rest, the water level in d will be at the same height 
as in a and stand below the overflow openings c'. But on the suction stroke of 
the engine, the vacuum produced in the vaporizer chamber a of the producer, 
Fig. 424, communicates itself through e to the interior of the tube c. This causes 
the water column in d to rise, some of the water o ver flowing through c' and so into 
the vaporizer. The quantity of water delivered increases with the suction, and 
is hence adjusted to the load on the engine, keeping the hydrogen content of 
the gas always at the most favorable point. The quantity of water delivered 
may also be regulated by hand by turning the cap / so that some air will be 
admitted through groove /', thus partially neutralizing the suction. 
































































































































































































































AUXILIARIES 


281 



Fig. 429 —General View of a 50 H.P. Guldner 
Suction Gas Plant. 

(Producer, scrubber, dry purifier, and tar 
extractor.) 


Pipe 

Connection 
to Suction 
Main of 
Engine 


I To Gas 
’Producer 

Fig. 430. —Water Supply Valve, Schweiz. Locomotiv- und 
Maschinenfabrilc, Winterthur. 

Piston b, fitting lightly in cylinder a, is raised every time 
the suction of the engine creates a partial vacuum in the 
gas main and in the space a. This action raises the needle 
valve c, which then momentarily opens the outlet d to the 
producer. The regulating screw e serves to adjust the 
tension of the spring and thus controls the lift of b. 


Fig. 431.—80 H.P. Capitaine Suction 
Gas Plant with Spray Scrubber and 
Centrifugal Washer. 1 

Tube vaporizer, shaking grate, grate 
bars of trough pattern filled with ash 
for protection against heat. Prelimi¬ 
nary cooling and washing of the gas in 
a scrubber with water supply at the 
bottom. Main scrubbing in a centri¬ 
fugal washer operated by the engine at 
a speed of about 100 r.p.m. Centrifu¬ 
gal action is supposed to take out tar 
and traces of sulphur as well as dust 
and water. 



1 Lecture given before the VI Regular Convention of the Schiffbautech. Gesellschaft. 























































































































































































































■282 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


(c) Gas Producer Installations for Fuels Carrying Tar 


About the only grade of tar-carrying (bituminous) fuels that is used to any 
extent to-day for the manufacture of power gas is the briquetted soft coal 1 or 
lignite. These briquettes, on account of their uniform size and their comparatively 
low content of water and ash, are fully as suitable for this purpose as anthracite 
or coke, provided of course that the tar is taken care of. If the tar is removed by 
means of washing and scrubbing apparatus outside of the generator, it is even possi¬ 
ble to use the ordinary type of producer. This subsequent separation of tar from 
the gas, however, soon causes operative difficulties with the ordinary type of scrubbers, 
and, outside of the fact that the rotary washers require power and that the odor 
of the scrubber water is in many instances very obnoxious, the method possesses the 
disadvantage, from the economic standpoint, of wasting the heat contained in the 
tar-forming hydrocarbons. The existing special constructions of soft coal (brown 
coal) gas producer installation therefore, in general, all endeavor to fix the tarry gases 
in the producer itself. The general method of doing this is to collect the gases of 
distillation and to lead them up through the incandescent bed of fuel above the 
grate, either simply breaking up the tar or burning it by supplying air, in which 
case the heat of combustion enters the general gasification process. This method of 
operation has been comparatively successful in properly designed installations; it is, 
however, still beset with many practical difficulties. The starting of such generators 
is a rather complicated process, for which reason, rather than iet them burn down, 
they are banked even for periods covering several days. During operation such pro¬ 
ducers under certain conditions exhibit a troublesome variation in the position 
(traveling) of the incandescent zone, in which case the tarry' gases are apt to escape 
undecomposed. 

In the design of brown coal or lignite generators, the generally very high per¬ 
centage of water in this kind of fuel should not be left out of account. This may 
be so high that the generator requires no further water supply during operation. 

This does not mean, of course, that the vaporizer may be entirely dispensed with 
in the common types, but the fact calls for wide range of possible adjustment in 

the quantity of steam made or that introduced to the generator. 

Soft coal gas producers are only in a few cases purely suction gas generators. 
In general the friction losses and resistances in generator and washing apparatus are 
too great to operate the installation on this principle, and fans are used to draw 
the gas through the producer and to force it to the engine. It is quite usual 

in this construction to make the fan act also as a part of the washing apparatus by 
the introduction of a spray of water which is afterward drawn off from the lowest 
point of the fan housing together with the dust picked up.2 Bituminous coal power 
gas installations, less simple in their construction than anthracite or coke plants, 
also cost somewhat more in regard to maintenance and attendance than the latter 
and their use consequently offers but little advantage for the smaller installations. 


1 Translator’s Note In this country, briquetted coal is as yet not used to any extent; where 
sott coal is used it is usually in the raw state. 

2 Fan washers have been frequently used for a number of years, as, for instance, in the operation 
of gas engines with blast-furnace gas, consequently their use can no longer be restricted by any 
patent rights. Recently, however, Theissen has claimed that the use of blowers working in this 

t water in J ection ‘institutes an infringement of his patent rights, and makes their further 
:r J Ct l +° h 1 ‘ S . llCei | 1Se - Thls contr oversy has an important bearing upon the gas-engine industry in 
general, and to help clear matters the writer has just caused the Gasmotorentechnik (April, 1905), to 
start a public discussion of the question. v " 



AUXILIARIES 


283 



Fig. 432.—Double Zone Generator for Brown Coal, Gasmotoren-Fabrik Deutz. 

This type of generator, in addition to the combustion zone just above the grate, has another at about 
the level of the vaporizer, near which is also located the gas exit. The fresh fuel fired on top of the upper 
zone is first subjected to distillation. The gases formed are drawn downward through the incandescent layer 
of coked fuel by the action of the exhauster connected in beyond the scrubber. This serves to convert these 
hydrocarbon gases into permanent gases. Air is drawn at the same time through openings in the coaer, thus 
causing a partial gasification of the coked fuel near the top, which action furnishes the heat necessary for dis¬ 
tillation and coking. The coked fuel slowly sinks downward, not all being gasified in the upper zone, and thus 
maintains the combustion zone above the grate. The action of the exhauster causes air to enter the vaporizer 
and the air-steam mixture so formed flows in under the grate and through the lower zone resulting in the pro¬ 
duction of ordinary producer gas. The cleaning apparatus consists of a wet scrubber, a centrifugal exhauster 
with water injection and a settling chamber. The bell of the gas holder controls a by-pass valve on the 
exhauster and thus regulates the draft on the producer. 



Fig. 433.—Brown Coal Gas Producer Installation, 30 H.P., Korting Bros. 

Besides the ordinary flat grate there is furnished a second inclined grate at the side, to aid in starting 
and cleaning the fire. The operation is otherwise much the same as for Fig. 432 the gas bei^g taken o 
. u Vioif wiv nr) instead of at the top The tarry hydrocarbon vapors formed at the top are sucked back 
toimMhe “By coked kyers of fuel and, meeting the current of air and steam coming ■up from below, 
are either burned or fixed. A special vaporizer is not required in this case on account of the high mois r. 
content of the fuel. The blower at the left is used at starting. 
































































































































































284 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 



Fig. 434.—Deschamps Producer for Brown Coal and Peat. 

In this producer all of the gas made is drawn out at the bottom. 
The producer proper, a, carries a fuel hopper or magazine 6, in which 
is suspended a cast iron bell c whose lower end is located in the upper 
part of the incandescent zone. The gases of distillation gather in c, 
but they are drawn downward together with the pre-heated air, or air- 
steam mixture, coming from the pre-heater and vaporizer e, by means 
of a fan connected to the outlet g. In passing the gas heats e. The 
vaporizer / furnishes a certain amount of steam to the air as it enters 
the pre-heater coil, in case dry fuels are being gasified. Grate bars, d 
may be shaken independently of each other and a poke bar inserted 
through the ash pit door serves to help keep the grate clean. For 
arrangement of entire plant, see Figs. 435-437 following. 



Figs. 435-437.-General Arrangement of a Producer Installation for Fuels forming Tar, Deschamps Type. 





















































































































































AUXILIARIES 


285 



Fig. 438.—Double Producer, Emil Capitaine. 


The fuel is forced in by a screw from below and be¬ 
tween the two grates. The fresh fuel is coked and the 
gases of distillation, in passing through the incandescent 
zone, are either burned or fixed. The mechanical feeding 
of the coal is also intended to prevent coking or clinker- 
ing. 


Details of Design and Dimensions of Gas 

Producers. The principal dimensions of pro¬ 
ducers, and especially of the combustion space 
proper, do not depend upon any theoretical 
considerations, but are based entirely upon 
experience and the results of practical tests. 
Fundamentally considered, the gas producer 
is a closed furnace in which a certain amount 
of coal must be burned on a given area of 
grate. How large the grate must be for that 
purpose depends, as in the case of other 
furnaces, directly upon the draft, the method 
of introducing the air, the kind of fuel, the 
depth of the fuel bed and what might be 
called the ‘Toad factor” of the installation. 
It is clear, however, at the outset that the 
grate and combustion chamber of a gas pro¬ 
ducer may be forced to much higher capacity 
than is possible in an ordinary furnace, because 
in the former the fuel is not completely 
burned but only gasified, and mostly with 
artificially increased draft. 



Section A-B. 


Figs. 439 and 440.—Brown Coal Gas 
Producer, Crossley Bros. 

The drying and coking of the green 
coal takes place in the three retorts a 
which are located in the upper part b 
of the producer and are heated by the 
gas leaving. The gases of distillation, 
through d, e, f } and g, are led under 
the grate, which may be shaken by 
turning it about its axis. 'Ihe air, 
which has been preheated, reaches the 
interior through the annular space h, 
as shown by the arrow. The retorts 
are separated from each other by 
slides i, and are charged and dis¬ 
charged in rotation. The coke is 
discharged by dropping the valve and 
scre.v k and then turning it about its 
axis. 























































































































































286 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Marine boilers with forced draft burn in the neighborhood of 30 lbs. of coal 
per sq.ft, of grate, locomotive boilers sometimes show as high as 100 lbs. per sq.ft. 
The former quantity of coal would in gas producers certainly yield 30X55 = 1650 
cu.ft. of producer gas. Assuming that an engine uses 100 cu.ft. of this gas per 
B.H.P. hour, each sq.ft, of grate or cross-section of producer would supply 16.5 H.P., 

or conversely, each 100 H.P. would require in round numbers 6.0 sq.ft, of grate or 

producer cross-section. In practice similar figures are quite often found in smaller 
suction gas plants. With the ordinary sizes of anthracite coal, however, the com¬ 
bustion may be forced somewhat beyond this point, up to about 40 lbs. per sq.ft, 

of grate per hour, and larger installations are often operated at this figure. 1 * * * 

The average cross-section of grate or combustion chamber should therefore be 


for every normal H.P. 


8.5 to 7 sq.in. for installations up to 25 H.P. 
7 to 5.5 “ “ “ above 25 H.P. 


( 1 ) 


Very small fuels, as for instance buckwheat or rice coal, require a larger 
cross-section in order to obtain sufficient area for the passage of air and to be able 
to work with a thinner fuel bed. Fuels low in heating value also call for increased 
areas. 

The internal height of the producer depends in the first place upon the heat 
consumption of the engine, the heating value of unit volume of the fuel, the time 
of combustion of one producer charge and finally also upon the behavior of the 
incandescent layers of fuel. Assuming that the heat consumed by the engine on the 
basis of fuel (that is, including generator losses) is 16,000 B.T.U. per B.H.P. hour, 
and that each charge of the producer lasts from 3 to 4 hours, the volume of the 
producer must be such as to store from 50,000-64,000 B.T.U. for every horse-power 
developed, or on the average 60,000 B.T.U. Now the heating values of some of the 
fuels based upon unit volume are about as follows: 


1 Translator’s Note. These figures are evidently somewhat too high for American practice and 
they may be plainly labeled as applying to Continental practice only. The entire discussion following 
should be read with that fact in mind. As far as the writer’s experience goes, gasification capacities 
on this side of the water range from 10 to 20 lbs. per sq.ft, of grate per hour, with the average 
perhaps at 15 lbs. He has seen only one manufacturer of gas producers claim more than that, up to 
32 lbs. Actual data from tests in this country is as yet rather limited. Wyer in his Producer Gas 
and Gas Producers, says, p. 233: 

“The rate of gasification -in a gas producer is relative to the character of the coal' used. The 
best rate determined by experience for a pressure type producer is 12 lbs. of coal per square feet of 
gra e area per hour, although some makers have advised as high as 20 lbs. of coal. However, the 
exact limit of coal consumption is not known, as it is dependent upon a large number of empirical 
factors. Experience has also demonstrated that too rapid driving opens a wide door for the admission 
of adverse gasifying conditions.” 

As an example from English practice, Sexton, Producer Gas, p. 114, states that a Duff 8X10 
pro ucer wi gasify about 1000 lbs. of coal per hour, giving a gasification of about 14 lbs. per sq.ft, of 

th e “ i° U1 TL a 1 f gGr am ° Unt Per Square foot of S rate area ' But even assuming that 

28 lbs P16S ° n ^ of the cross-section, this would mean a maximum gasification rate of 

Guldne/i^nl n ° ted ! hat f b ° th ° f theSe exam P les are fr «m pressure producer practice, while 

for the latte^ ttn 0 e V % SP ° f SUC « 0n ^ as P roducer8 ‘ The writer unfortunately has no data on hand 

from our own hard^ ° ] pr ° ducer - Tke kmd of coal called anthracite by Guldner is not far different 
trom our own hard coal, containing about 85% C and little ash. 



AUXILIARIES 


287 


1 cu.ft. = 50 lbs. of anthracite coal contains, deducting for losses, 565000 B.T.U.; 

1 “ = 25 lbs. of coke contains, deducting for losses, 340000 B.T.U.; 

1 “ =50 lbs. of brown coal contains, deducting for losses, 340000 B.T.U. 

Hence, for a charging period of from 3 to 4 hours, the generator volume should be 
for operation with anthracite coal at least .105 cu.ft. j 

“ “ “ coke “ “ .175 “ for every D.H.P. . (2) 

“ “ “ lignite “ “ .175 “ J 

The vertical height required for similar fuels increases with the size of such 
fuel. For very small sized fuel the fuel bed should be kept thin in order ta 
reduce frictional resistance; the only way to get sufficient space in such a case is 
to increase the diameter of the producer correspondingly. On the other hand, a 
fuel bed that is too thin interferes with proper gasification, allowing both air and 
carbon dioxide to reach the upper generator space undecomposed, thus decreasing the 
volume of the gas and causing some of it to burn in the producer itself. It is 
better, therefore, to make the fuel bed too thick rather than too thin. This is of 
special importance for a coking coal, which, due to fus¬ 
ing together, may open up wide passages through the 
bed, thus seriously affecting the gasification process. 

Finally, an ample fuel space in the generator is one of 
the best assets where the fluctuations in the load are 
apt to be strong and long continued. 

The construction of the generator shell and lining is 
the most important thing, determining the satisfactory 
operation of the producer as well as the value and life 
of the entire installation. In spite of this fact, far too 

little attention is paid to-day, even in their design, to 

the experiences gained in the building of industrial gas 
producers (Siemens type, etc.), and especially blast fur¬ 
naces. There are suction-gas generators built in which 
the lining is constructed entirely at variance with the 
principles of gas technology, and the renewal or repair 
of which would call for the labor of skilled hands for 
days. Even the life of properly constructed linings is 
comparatively short; under the best conditions they may last about two years, 
under adverse circumstances often only a few months. Ease of repair and of renewal 
is therefore the thing to be provided for under any circumstances. Single-piece 

linings are not satisfactory even for small generators, for large producers they are 

absolutely unfit because they crack easily under heat and become useless, while their 
renewal is always accompanied by considerable expense. 

The best way is to utilize in the first place a fire brick (not too fine grained 

or too hard burned on account of cracking) w’hose quality has already been tested 

in service, and to construct the lining of several rings joined as shown in Figs. 441 

and 442. Each ring is divided into from 4 to 10 segments (or radial brick for the 

larger sizes). The lower ring wears out the fastest, and for that reason arrangements- 
should be made to enable the renewal of this by itself with ease. 

It is poor design to support the lining on a ring of metal solidly riveted to the 
generator shell. It is better to support it on a separate and divided ring of cast 



Figs. 441 and 442. —Sectional 
Producer Lining. 


























288 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

iron, whose various parts may be easily replaced through the ash-pit doors when 
the side next the fire has been burned out or melted away. (See Fig. 426.) 

The flat grate so generally used in producers for the medium sizes of anthracite 
and for coke has air spaces varying in width from T V' to tV*- Too wide air spaces 
allow too much of the fuel to fall through, while those too narrow soon clog up 
with ash and clinker, thus interfering with the cleaning of the grate. Very small- 
sized fuel requires the use of step or basket grates; while, some other fuels, as yet 
little used, also call for special grates. Still others, like wood and charcoal, may be 
gasified in special generators without grates. 

As determined from practical experience, the starting fan in installations up to 
50 H.P. should have a capacity of about 175 cu.ft. of free air per minute. Larger 
installations require capacities of from 350 to 700 cu.ft. per minute, depending upon 
horse power. On account of low cost it is usual to employ hand-operated fans in 
which hand-crank and multiplying gear form a part of the outfit. It is better, 
however, in every case, and in large installations indispensible, to have the fan 
operated by a special source of power (small water wheel, electric motor, etc.) in 
order to leave the attendant free for his other duties during the period of heating up. 

2. Gas Washers and Purifiers. The gas as it leaves the producer carries along 
with it a certain amount of fine dust which must next be removed. Depending upon 
the kind of coal used, the gas also contains more or less chemical impurities, such as 
tar, sulphur vapor etc., each of which may cause trouble in operation, and should 
therefore be removed as far as possible. The common types of wet scrubbers are, 
if properly constructed, usually able to successfully throw down most of the mechan¬ 
ical impurities in the power gas. In spite of this washing, however, there is usually 
still enough fine dust left in the gas to cause scoring of the internal rubbing surfaces 
of the engine. As a matter of safety, therefore, it is always best to pass the gas 
through a dry purifier after it passes the scrubber. 

The chemical impurities, when they occur in any considerable quantity, cannot 
be successfully eliminated by any of the common washing apparatus. Neither hard 
coal nor coke cause any trouble in this respect, because the quantity of tar found 
in the gas is so small that it can be easily handled with the simple means in common 

use. The latter, however, completely fail with gas from bituminous coal, for which 

reason alone the tar-forming constituents of this gas should be changed to fixed 
gases in the producer itself, i.e., they should not be allowed to reach the cleaning 
apparatus. An exception to this rule is found in some of the very large industrial 
gas-generating plants, in which special apparatus is used to separate the tar and the 
other impurities and to obtain them as by-products. 

Wet purifiers (washers, scrubbers) usually consist of tall cylindrical vessels through 
which the gas passes from bottom to top, being washed by means of fine streams 
of water which it meets in its ascent. In order to promote a thorough contact 
between gas and water, the purifiers are usually filled with coarse coke, sometimes 
also with wooden baffles, etc. There is a free space at the top to allow the water 

to spread evenly, and one at the bottom to serve as a settling chamber for water 

and mud. The chamber should be provided with ample exit and cleaning openings. 

The water-swept space of the scrubber should be at least .7 cu.ft., better still 
from .9 to 1.1 cu.ft., for every engine horse-power. The height should be as great 
as convenient, in order to provide long paths for water and gas. Purifiers of low 
height but large diameter should be used only when the water is very thoroughly 
sprayed and distributed. 


AUXILIARIES 


289 


The water may be finely divided by any of the well-known spray apparatus 
or by baffle plates. Regarding the use of the former, it should be noted that fine 
openings may be stopped up internally by impurities in the water and externally 
by tarry deposits, in either case causing trouble. For that reason any spray appar¬ 
atus used should be so arranged that it can be easily taken out and cleaned. The 
dirty water should be taken out at the bottom of the washer by a pipe of ample size 
into a syphon or seal box. The latter should be tightly closed at the top to 
confine the very obnoxious odor of the water, and further should be located at least 
12" lower than the scrubber to prevent a sucking back of the water. 

The so-called centrifugal washers, if properly made, both clean and cool the gas 
very thoroughly. The simplest form of this washer consists of a fan (exhaust fan 
or blower) in which water is sprayed into the casing (see Figs. 432 and 435). There 
are, however, special centrifugal washers 1 on the market. (See Fig. 431.) The 
latter are very well adapted to large installations and have been used for some time 
up to the very largest sizes (in blast-furnace gas installations for example). The 
thing that prevents their general adoption in commercial gas producer installations 
is the fact that they require a power drive, which, in most cases, is not available 
in the producer room. 

Dry purifiers are cylindrical or box-like vessels which are furnished with from 
2 to 4 horizontal perforated partitions which serve to support the purifying material. 
The latter may consist of wood shavings or sawdust, excelsior, chips of iron or steel, 
slag wool, etc. Any material that contains fine, sharp particles, or which disinte¬ 
grates under the action of the gas and thereby forms fine and hard particles which 
the gas may pick up, is unfit for use. Slag wool and iron chips may be specially 
harmful in this direction. The free passage for the gas should be the greater the 
finer the purifying agent; sawdust, for 'that reason, requires the most roomy dry 
purifiers. On the average, the approximate volume of dry purifiers should be about 
.2 cu.ft. per engine horse-power; but, where possible, from .3 to .5 cu.ft. should 

be allowed. 

Several firms also connect into the gas mains just beyond the scrubbers a tar 
separator or extractor. These are generally based on the principle of Pelouze, 

according to which tar vapor when made to strike metallic plates at right angles 

under high velocity, will condense and gather on these plates in drops. See Fig. 429. 
Care must be taken to see that these baffle walls can be exchanged quickly after 
having been coated with tar. The best-known, but at the same time least desirable, 
form of this apparatus consists of a cylindrical wire brush, which is very often placed 
in the gas main close to the engine. These brushes, even when clean, increase the 
friction resistance in the main considerably, and becoming dirty after only a few 
days of operation, so seriously interfere with the suction that the normal load of 

the engine can no longer be attained. _ 

The path for the dirty gas, that is, the piping between producer and the first 
scrubber, should be as short as possible and so arranged as to offer the least chance 
for the lodgment of any soot or ashes. Whenever these are deposited, however, 
ample opportunity for cleaning in the simplest and quickest way should be provided. 
This requirement is hard to fulfil if, as is usual, the vaporizer or pre-heater is built 
into the gas main between generator and scrubber. It is equally out of place to put 
the dry purifier close to the engine, in which case the still partially uncleaned gas 


1 For example, that made by Edward Theissen, Munich. See footnote, p. 282. 



290 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


is forced to pass through considerable lengths of pipe, giving it a chance to clog 
them up. All cleaning apparatus belongs close to the producer and should not be 
placed in close proximity to the engine. 

The gas receivers often used between the cleaning apparatus and the engine are 
of little value for purifying the gas. They may act as gas holders or pressure 
regulators only when their volume is made much greater than is usually the case. 
This volume, if the receiver is to serve any purpose at all, should never be less than 
ten times the piston displacement. 

It should not be forgotten to supply the pipe line just in front of the engine 
with a suitable arrangement through which the system can be cleared of air and 
poor gas. The purge pipe should not be combined in any way with the try-cock, 
used to try the quality of the gas, but should be a line leading directly outside to 
avoid fouling the air in the producer room. The try-cock should be opened only 
after the poor gas has been cleared out. 

Last, but not least, the designer should, besides taking into account technical 
and practical requirements, also heed whatever government or insurance specifica¬ 
tions may exist in his locality regarding the construction and erection of gas pro¬ 
ducer installations. Such supervision does not as yet exist in many places, but the 
designer will do well to take the most strict of those in in existence, as a guide to 
avoid encountering difficulty in the sale of the apparatus in any territory. 


II. Starting Apparatus 

An internal-combustion engine is not self-starting, but must by some suitable 

external means be set in sufficiently rapid motion before it can do the work required 
to overcome its internal resistance. If the starting is done by “cranking” or turning 
the shaft, the starting agent may continue its action even after the first ignitions 

have taken place, in which case both sources of power combine to start the engine. 
If, on the other hand, the source of power for starting is made to act in the cylinder, 
it is usual to impart to the flywheel only sufficient velocity so that the inertia of 
the mass may be enough to overcome the starting resistance. After this velocity is 
attained, the starting agent is shut off, the first ignition takes place in the cylinder 
and soon brings the engine up to normal speed. To reduce the starting resistance 

it is usual to reduce the compression on starting as far as the kind of fuel used 

will allow, for rich fuels to from 1-2 atm., for lean fuels to from 3-4 atm. This 

is done by forcing a part of the charge through the exhaust valve, i.e., not closing 

the latter until some time into the compression stroke. It is also common, at least 
in large engines, to so retard the spark that ignition cannot occur until after the 
dead center is passed in order tb avoid the back-firing so dangerous to attendants. 

The usual methods of starting serve their purpose only when the engine is light, 
but will not start an engine under load. Only twin-cylinder engines may be started, 
at least under partial load, when one cylinder is operated with compressed air while 
the other operates on fuel. But this scheme requires large air storage-tanks and is 
even then not quite certain. 

A suitable starting gear should in all cases be a part of an up-to-date engine 

installation. For large engines it is of course a necessity, but even for the smaller 

units it is a good thing from the standpoint of safety. The starting of an engine 
by turning the wheel is dangerous for unskilled hands, and German trade-unions, at 


AUXILIARIES 


291 


least of late, therefore require that even small machines shall be furnished with 
proper starting apparatus. 

1. Hand Cranks. It is possible to start engines up to 10 H.P. by means of 
hand cranks. Larger engines, up to say 20 H.P., may be started in a similar way 
if the motion is transmitted through 
gears or sprockets and chain, as 
shown in Figs. 443 and 444. In 
such a case it is best to arrange 
the drive so that the operator is 
able to bring his greatest power 
to bear on the crank at the mo¬ 
ment of greatest resistance in the 
engine, i.e., near the end of the 
compression stroke. In four-cycle 
engines this means that the trans¬ 
mission ratio should be either 2:1 
or 4:1. 

The types of hand cranks a 
and 6, shown in Figs. 445-447, 
automatically lose their grip on 
the shaft as soon as the first normal 
explosion occurs, but they remain 
in mesh if through back-firing or 
pre-ignition the shaft should start 
in the opposite direction. This 
may, under the circumstances, be 
dangerous to the attendants and 
may be avoided by the use of 
safety cranks which free themselves 
from the shaft when it starts in 
either direction of rotation. An 
example of this starting crank is 
that made by Dr. W. Heffter of 
Berlin, N.W., in which, when the 
shaft starts in the normal direction, 
an axial bolt simply unscrews 

out of the end of the shaft, while for rotation in the opposite direction the con¬ 
nection is broken by a ratchet arrangement. This safety crank is made for engines 

up to 16 H.P.; for larger sizes gear or chain drives are 
used in connection with it. 

2. Mechanical Starting Apparatus. Where a second 
engine is available, it, or the transmission it operates, may 
be used to start the engine. Special arrangements for 
this purpose are usually not necessary. The scheme for¬ 
merly used, of supplying each larger machine with its own 
starting engine, is seldom employed to-day. Where water 
under high pressure can be had, the use of a hydraulic piston motor of some kind 
might be considered, especially if the water used can afterwards find application in 
the jackets of the main engine. 



Figs. 443 and 444.—Starting apparatus with safety arrangement 
against back-firing, 15 H.P. Giildner Engine. 



Figs. 445-447.—Starting Cranks. 

























































292 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


3. Starting by Means of the Fuel Mixture. The general idea of this method is 

to place the engine crank 10 or 15% beyond head dead center, to fill the combus¬ 
tion chamber with the fuel mixture, and to cause ignition by snapping the spark by 
hand. The piston thus receives an impulse which is usually sufficient to impart 
enough energy to the flywheel to last for several further turns, during which the 

normal charge is drawn in, compressed and ignited. There are several methods of 

carrying out this scheme, which differ among themselves in the mechanical means 
employed. Clerk caused the fuel pump of his 2-cycle engine to pump the normal 
fuel mixture into a reservoir during the time that the igniter was shut off and the 
engine slowing down. The compressed mixture was first led into the power and 

pump cylinders and acted as the direct source of power to cause the first rotation. 
At the end of expansion the mixture was not allowed to escape, but was compressed 
on the next instroke and ignited. Lanchester utilizes the sucking action of the 
exhaust pipe to draw a combustible mixture into a vessel connected to the com¬ 
bustion chamber during the time that the exhaust valve was open. This mixture 
is at the desired time exploded by means of an open flame in connection with the 
above-mentioned vessel. The gases of combustion immediately flow over into the 
combustion chamber and give an impulse to the crank which has previously been 
set at the proper starting position. This arrangement was afterward improved by 
connecting gas and ignition cocks directly to the cylinder. The latter, during the 

slowing down of the engine, has been thoroughly scavenged and filled with clean 
air. On starting, the gas cock at the side of the cylinder is opened. A second 
cock on top of the cylinder allows of the escape first of some of the air displaced 
by the gas, and soon thereafter of the mixture formed. The latter is ignited by an 
open flame in front of the cock. As soon as the. needle flame formed shows that 
the composition of the mixture is right, the gas cock at the side is closed. This 
stops the flow of mixture out through the ignition cock, the flame there burning 
strikes back into the interior of the cylinder and explodes the mixture still remain¬ 
ing. The pressure so generated closes a small check valve in the ignition cock and 
gives an impulse to the piston. The smaller Simplex (Cockerill) engines are started 
in a similar way. The large blast furnace gas engines of the same firm are put in 
operation by pumping a gasoline fuel mixture behind the piston when in proper 
position, compressing this mixture slightly and exploding it by means of electric 
spark. The scheme used by the Berlin-Anhaltische Maschinenbau Gesellschaft for the 
starting of their large engines, consists in allowing a small gas engine to compress 
a part of its fuel mixture through a third valve into the combustion chamber of 
the large machines. During this operation the piston of the latter is held in the 
starting position by blocking the fly-wheel. The mixture is exploded by electric spark, 
the impulse breaks the cast-iron block at the wheel, freeing the shaft. 

Starting by mixture is a simple process in the case of illuminating-gas engines. 
For other fuels it is somewhat complicated and not absolutely reliable. If the engine 
should fail to start at the first trial, which may occur even with the best of machines, 
there is often not enough mixture in reserve for a second attempt, and it becomes 
necessary to go through all of the operation again, which often takes considerable 
time. Especially in the case of large engines, the work of refilling the reservoir or 
cylinder with mixture is generally troublesome and tedious. 

4. Starting by Compressed Air. This method, which enjoys general preference 
to-day, was already in use in the old constant pressure 2-cycle engines of Brayton 
and Simon, in which air in receivers under a pressure of from 60-75 lbs. was con- 


AUXILIARIES 


293 



Fig. 448.—Combined starting and 
back pressure (or back flow) 
valve, 80 H.P. Horizontal En¬ 
gine. 

(Valve is operated by hand while 
starting.) 


stantly available. The pressure of the scavenging air used in our modern 2-cycle 

explosion engines is not high enough for starting. The starting air is usually com¬ 

pressed either in the main cylinder or by means of an independent small compressor. 
According to the former scheme the piston of the main engine during the slowing 

down period, after the fuel is shut off, compresses the air drawn in into a special 

receiver through a valve, which is generally 
used also as the starting valve. The re¬ 
ceiver should be of such size that the 
engine can compress the air up to 75 or 
90 lbs. With this pressure it becomes 
• necessary to give the piston several impulses 
on starting. On the other hand, where 
an independent compressor is used, it is 
advisable to compress the 
starting air up to from 150 
to 225 lbs., and to accomplish 
the starting with few, but 
powerful impulses. With an 
installation properly arranged 
and adjusted (good gas, 
proper mixture, correct spark 
timing, right position of 
piston, etc.), a single impulse 

is generally sufficient to start with certainty engines, with ordinary fly-wheels, up to 
150 H.P. Especially where the starting valve is operated by hand, it is of im¬ 
portance to start with one or at most two impulses, because the proper handling 
of the valve becomes difficult as soon as the engine speeds up. 

If the valve through which the main piston delivers air 
into the receiver is also used as the starting valve, its valve 
disk must open outward, and must, during normal operation, be 
held against its seat by means of a screw, see Fig. 448. Such 
a valve, however, will leak easier than if the disk opened 




4-k f 





Fig. 449. —Starting valve, 
Hornsby & Sons. 

(For location of this valve, 
see Figs. 240-242, p. 197. 
Fig. 450 shows the method 
of operating it.) 



Fig. 450.—Starting arrangement for a 40-50 H.P. Hornsby-Akroyd 

Oil Engine. 

[Operated by hand so that air is admitted at the beginning of the third 
stroke (expansion stroke)]. 


UI upci O/Clllg IKI'J - - - a 

For starting, the lever is placed in the position a, Fig. 450. , Air compressed to 75 or 90 r^wh pti° si o n^'down'f or a'stop^ 
and a, Fig 449, into the cylinder. The air is obtained byletting.the engine act as a comp ■ , In the middle position 

For this the lever is put in the position 6, which raises valve b while a acts as a back p - •. , ^ which is the 

c of the lever, Fig. 450, the internal lever c, Fig. 449, leaves both valves free, a is then held down on its seat by d, wmcn 

normal operating position. 






































































































294 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


inward and were held against its seat by the force of the explosion. A further dis¬ 
advantage of the “slowing down” method of obtaining compressed air for starting 





Air Tank, 


Y///////////y/y//////////// 



i - j| 


- A -1 


Figs. 451 and 452.—Starting Air Compressor for Engines from 60-80 H P. 



consists in the trouble of getting 
air for starting the first time 
and the refilling of the receivers 
when the charge of air is sud¬ 
denly lost through leaks. 

For large engine installa¬ 
tions preference should be given 
to independent air compressors 
driven by their own source of 
power wherever possible. Any 
connection or interdependence 
of air pump and main engine 
is bad, because in case of re¬ 
peated failure to start, the air 
will be used up without any 
chance of renewal by means of 
the engine. This may lead to 
serious trouble. Under such cir¬ 
cumstances, recourse has already 
been had to compressed carbonic- 
acid gas, which can be bought in 


steel flasks in the market. But this scheme is not entirely sure, since the residual carbon 
dioxide makes the mixture poor and hard to ignite. ' This latter fact has probably 








































































































AUXILIARIES 


295 


caused the Gasmotorenfabrik Deutz to abandon its patented method of starting with 
compressed exhaust gas. (This method consisted in allowing a small part of the 
burned gases at each explosion to flow through a small heavily loaded valve into a 
receiver from which they were afterward used to start the engine. The valve men¬ 
tioned was entirely closed when the receiver pressure indicated from 100 to 120 lbs.) 

An example of a small starting compressor is shown in Figs. 451 and 452. This 
is designed to compress air up to 180-225 lbs., and requires from 1.5 to 2.0 H.P. 
to operate it. Since the compressor is only used a few moments every day, the cooling 
of the compressor cylinder may be dispensed with. The machine is normally operated 
by belt, but the wheel is also fitted with a crank, which may be used as a make¬ 
shift. The arrangement of a compressor of this type with its receiver may be seen 
from Fig. 453.’ The volume of the tank should not be less than .35 cu.ft. for each 
normal engine horse-power, i.e., about 10 times the piston displacement. In important 
installations the use of two such tanks is advisable, both being kept under pressure 
and the second used as a reserve for the first. 

5. Starting by Electricity. Gas engines which are used to drive electric generators 
may be started, in case a storage battery or another source of current is available, 
by operating the dynamo for a few turns as a motor. The use of current in this 
manner, of course, assumes that there is sufficient capacity in the battery. The 
actual arrangement of the system of wiring used for the purpose may be seen from 
the diagram of connections, Fig. 454. This shows that the normal wiring system is 
increased only by the addition of a double-throw switch and a connection shown in 
broken line in the figure. The method of operation 
. is very simple, it is merely necessary to close a double¬ 
throw switch, to regulate the position of the charging or 
discharging levers and to adjust a resistance box. As 
soon as the engine picks up its cycle and the starting 
current drops to k certain point, an automatic circuit 
breaker acts, and the dynamo, which up to this point 
has acted as a motor, operates under no load until the 
proper switches are thrown and the circuit is closed. 

The use of a storage battery as the source of 
starting current, is only feasible for the smaller 
engines, say up to 50 H.P., and even for these the 
service is rather severe on the battery. It is quite 
possible to totally destroy small batteries in a few 
days when used for starting,- especially if the matter 
is not skilfully handled. The engine should always be 
started with the piston in position at the end of the 
compression stroke, in order that the electric motor during the next three strokes is 
required to accelerate only the fly-wheel and to have the kinetic energy of the latter 
aid in the next compression stroke. It is totally wrong, even with the largest and 
most powerful battery, to commence starting with the piston, for instance, at the 

end of the suction stroke, because in such case the motor must not only accelerate 
the moving parts, but in addition is called upon to overcome the maximum resist¬ 
ance at the crank. It is probably best in all cases to start by holding open the 

exhaust valve, and thus calling upon the electric motor only for enough energy to 
bring the fly-wheel up to the desired speed, and after this is accomplished to set 

the valve gear to normal position. 



Fig. 454. —Lay-out of Connections 
for usrng Generator as a Starting 
Motor. 





























296 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

The figures in the following table will give some idea of the power consumed 
in overcoming the starting resistance: 


Table 27 

POWER CONSUMPTION OF STARTING-MOTORS 


40 H.P. Deutz engine driving A. E. G. Dynamo, 

110 V., 200 Amp. 

60 H.P. Deutz engine driving A. E. G. Dynamo, 

110 V., 450 Amp. 

1. Second 

2. 

4. 

6. 

7. 

8. “ 

112 volts 
111 “ 

111 “ 

110 “ 

110 “ 

110 “ 

30 amp. 
40 “ 

45 “ 

60 “ 

90 “ 

110 “ 

/1st ignition 
\ engine starts 

1. Second 

2. 

3. “ 

4. 

5. “ 

6. 

7. “ 

110 volts 
110 “ 

110 “ 

110 “ 

110 “ 

110 “ 

110 “ 

50 amp. 
90 “ 

100 “ 

120 “ 

120 “ 

100 “ 

90 “ 

/1st ignition 
l engine starts 


According to an empirical formula, based by Leroy upon the results of several 
tests on the subject, the maximum power required to start by electricity in the 
manner described is approximately f of the maximum capacity of the dynamo 
concerned. 1 


III. Mufflers 

The noise caused by the air drawn into the engine and that due to the exhaust 
is annoying even in small engines; in large machines it may become a public nuisance. 
The various kinds of mufflers used to diminish or abolish this noise show a number 
of different constructions, but they all depend for their action either upon the gradual 
decrease of the air or gas velocity, or upon the principle of the Helmholtz “resonator,” 
in which the inlet or outlet pipes are enlarged into one or more equalizing chambers 
into which the sound waves expand with consequent reduction of intensity. The 
latter method is the one mostly used for exhaust mufflers. 

1. Inlet Mufflers. Besides decreasing the noise, inlet mufflers also serve the 
purpose of freeing the air from the coarser mechanical impurities and of any con¬ 
siderable amount of water that it may carry. Simple vessels, or receivers, used for 
this purpose are usually given a volume at least equal to five times the piston 
displacement, and where possible more. Baffle plates, perforated partitions, and 
similar means for reducing the noise make it possible to make the muffler volume 
considerably smaller. The wall thickness of the mufflers depends directly upon 
considerations of manufacture; the material mostly used, that is, cast iron, requires 
a minimum thickness of from £ to f”. Sheet-steel mufflers are most often used 
in combination with baffle plates, partitions, etc.; without these such mufflers require 


1 Revue industrielle, 1898, p. 229, according to L’industrie electrique, 1898. 














AUXILIARIES 


297 


considerable wall thickness in order to prevent resonant vibrations in the muffler 
walls themselves. 

It is possible, in many cases, to sufficiently deaden the noise by taking the air 
from the base of the engine, or by shaping the end of the pipe so that the inlet 
cross-section is divided into a number of narrow slots (.04 to .08") of ample length. 
In the former case, the air inlet to the interior of the frame or bed should not be 
placed in the side next to the fly-wheel because of the greater quantity of dust 
there stirred up, which latter then easily reaches the cylinder. 

It is distinctly not advantageous, in the case of vertical engines, to draw the 
air from a closed crank case. Under such conditions the air is considerably pre¬ 
heated, is consequently less dense, and may further be contaminated by oil vapors. 
On the other hand, taking the air from the room itself has the advantage of aiding 
in the ventilation and further may serve to draw out of the room some of the oil 
vapor formed, a fact more or less pleasantly noticeable, especially in small installa¬ 
tions. 

In the case of large engines the complete abolishment of the inlet noise is 

possible only at considerable expense, and then only with means that in most cases 
require considerable room. The most common way of doing this is to construct 

special pits or underground suction chambers which are connected to the engine by 
means of masonry conduits and have their own inlet openings in the open air, where 
there is least dust. In some cases large engines have been allowed to draw air 
from interior chambers in their own foundations, but the scheme is permissible only 
when the chambers have been finished inside to guarantee absence of dust and have 
otherwise been so constructed that a transfer of the sound waves cannot take 
place. 

2. Exhaust Mufflers. Exhaust mufflers, in order to be fairly efficient as regards 
deadening of noise, require a volume equal to from 15 to 20 times the piston dis¬ 
placement. In general, however, the volume is made only from 6 to 8 times the 
piston displacement at the expense of efficiency. To fulfil the strictest requirements 

as regards muffling, it is not only better but generally also cheaper to connect two 

or three smaller mufflers in series. As in the case of inlet mufflers, the addition of 
any scheme to gradually increase the volume or to gradually change direction of flow 
by baffles, etc., enables the use of exhaust mufflers of smaller volume, but in the 
use of such means care should be taken to see that the back pressure on the engine 
is not too greatly increased. In general, the slower the flow of gas from the cylinder, 
i.e., the slower the opening of the exhaust valves or ports, the smaller may the 
muffler be made. For that reason, engines with a ring of exhaust ports, for instance, 
require larger mufflers than machines having exhaust valves. In multicylinder engines 
it is a good scheme to furnish each cylinder with a muffler, in order to prevent any 

possible interference of the outflowing gases. 

Cast-iron mvtjie pots having an internal diameter of D inches may be given a 

wall thickness of about 


£ = — + .25", but S should never be less than f" 
50 

Their height or length should be about 


( 1 ) 


H = 1.25 to 1.75 D, 


( 2 ) 




298 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


Above D = about 40", the use of a wall thickness according to Eq. (1) results in 
cumbersome and heavy constructions. It is consequently better above this size to 
use sheet steel, as, for instance, old boiler plate from f to f" thick. 

The inlet and outlet openings of these muffle pots should be separated from 
each other as far as possible, the one for instance should be placed diametrically or 
tangentially above the bottom, the other axially in the top or head, see Figs. 457-458. 
Just above the bottom there should be a drain-opening at least 1" in diameter. 

The sheet-steel mufflers bought in the market rarely form a completely successful 
substitute for the muffle pot. They are internally strongly corroded by the exhaust 
gases even when galvanized, and in most cases increase the back pressure. Further, 
if the steel is not of ample thickness and the surfaces not sufficiently stayed, each 
exhaust may cause noisy vibrations in the muffler itself. 

Large muffle pots sometimes cause considerable annoyance on account of radia¬ 
tion of heat. For this reason, in several of the large engines, a part of the cylinder¬ 
cooling water is sprayed into the muffler, or the exhaust line may be used as the exit 
for all of the jacket water. In such cases, however, the greatest care should be 
exercised to see that the water can, under no circumstances, flow back into the 
cylinder or be drawn back by the piston. From a practical standpoint this scheme 
of cooling the exhaust gases has the disadvantage that the pipe is very apt to rust 
through and must be very thoroughly drained in cold weather. Further, the exhaust 
is constantly accompanied by a cloud of steam, which makes impossible a visual 
judgment of the quality of the exhaust gases. 


Designs of Mufflers: 



Exhaust Muffler, 6 H.P. 
Engine. 


Muffler for a 12 H.P. 
Hornsby Engine. 


The quieting of the exhaust noise 
in large gas engines offers even greater 
constructive difficulties than the muffling 
of the inlet. The common muffle pots 
finally assume very large dimensions, 
and even then do not fully serve their 
purpose. The best scheme is in this 
case, also, the construction of under¬ 
ground masonry ducts into which the 
gases are allowed to enter after the 
greater part of the expansion has been 
completed in an iron muffle pot. Prob¬ 
ably the most complete deadening of 
the noise is obtained if, after the ex¬ 
haust gas has been allowed to expand 
into the ordinary muffler or into a duct, 
it is drawn out of the latter by means 
of an exhauster running in a water 
spray, and thus discharged into the 
open air in a steady stream. The method, 
however, is rather costly, since it re¬ 
quires a steady source of power, a 
constant supply of cooling water and is 
subject to frequent repairs. 




























AUXILIARIES 


299 



Figs. 459and460.—Exhaust Muffle 
for a 25 H.P. Engine. 

(The cover may be shifted so as 
to bring the flange opening for the 
outlet pipe in the desired position.) 



Figs. 461-464.—Steel Plate Exhaust Muffler, 80-100 H.P. 
Engine. 


IV. Cooling Arrangements 

Although the smaller sizes of automobile engines are often cooled by some method 
of air cooling, stationary engines are almost without exception cooled by water. The 
cooling at least extends to the combustion chamber and that part of the cylinder 
uncovered by the piston in its travel, but it is in many cases also extended to the 
other less highly heated parts. 

Experience shows that on the average one-third of the total heat supplied per 
effective horse-power hour is lost in the external cooling medium. With the liberal 
assumption that the consumption of heat per H.P. hour is 12 000 B.T.U., it will 
be seen that the cooling medium must be able to carry off 4000 B.T.U. per H.P. 
hour. With a temperature range of 90° (from 60 to 150°), this corresponds to a 
consumption of cooling water equal to 45 lbs. (5.5 gals.) per H.P. hour, assuming 
that the hot water is wasted. 

In the larger engines the heat consumption per H.P. is considerably smaller 
than above assumed. This has the effect of reducing the amount of cooling water 
required (to about 4.25 gals.) .especially since the larger the machine the smaller will 
be the proportionate heat loss to the jacket. On the other 
hand, certain features of construction and of operation of large 
engines make it usually necessary not to heat the jacket water 
above about 125°, which will again tend to increase the quantity 
used. In case the fuel mixture carries easily inflammable gases, 
like hydrogen, it is well, in order to be on the safe side, to 
estimate the amount of cooling water required per H.P.-hour at not 
less than from 5.5 to 8.0 gallons, based upon maximum capacity. 

It is of advantage to construct separate inlet and 
outlet lines for the jacket water to the most important 
engine parts requiring cooling (cylinder head and barrel, 
valve cages and stuffing boxes, exhaust valve, piston and piston 



Fig. 465.— Direct Cool¬ 
ing of a Horizontal 
Engine. 


rods, etc.), so that 




















































300 DESIGN AND CONSTRUCTION OF INTERNALrCOMBUSTION MOTORS 


the supply to each can be independently regulated. For the water-cooled pistons of 
large engines a separate line is indispensible since the water must be under a pressure 
of from 60 to 75 lbs. per sq.in., while for the other machine parts a head of from 
25 to 35 feet is usually quite sufficient. 

Wherever it is not possible to obtain sufficient cooling water, that is, where it is 
obtained only with difficulty, or where its cost is high, it becomes necessary to cool 
the hot water by some artificial means in order to use it over and over. In this 
way only the escaping vapor or steam must be replaced and the consumption of 
cooling water may be cut down .8 gallon per B.H.P. hour. The cooling arrange¬ 
ments for this purpose in small installations usually consist simply of tanks whose 
capacity or superficial surface is sufficient to radiate the heat taken up from the engine 
during a limited period of operation. 

A somewhat forced means sometimes used to reduce the consumption of cooling 
water, consists in not only heating the water in the cylinder jacket, but in actually 
boiling it in the jacket space. Thus each unit of water carries away not only the 
heat of the liquid but in addition also the latent heat of vaporization, which is 
approximately five times as great as the former. Several disadvantages connected 
with this scheme, however, make it generally undesirable for use in stationary engines. 
In small alcohol engines it may under certain circumstances give satisfactory service. 
Altman in Berlin was the first to use this method of cooling in his alcohol locomo¬ 
biles, but the idea is common property and had already been used by Lenoir in 
1860 in his first gas engines. 1 

To cool the jacket water, there may be used 

1. Cooling Tanks, Fig. 466. They may be used in installations up to 50 H.P., 
and are constructed of galvanized sheet steel ^ or thick. The bottom is either 

made of somewhat thicker stock or suit¬ 
ably stiffened. The height may be from 
1.75 to 2.5 times the diameter. The capa¬ 
city may be made approximately equal to 

V = 1.4 z N n cu.ft.(3) 

provided the engine, of N n B.H.P. normal 
capacity, is in operation z hours daily 
(z f 10). It is recommended, however, to 
use two or more tanks if the necessary 
capacity V exceeds 175 cu.ft. The flange 
for the outlet pipe is placed either in 
the bottom or in the side close to the 
bottom while the inlet opening is situated 
from 6 to 10" from the top. To best 
fulfil its purpose the tank should be placed 
in a cool, airy spot, but of course not too far from the engine. On account of the 
danger of freezing it is not safe to place it outdoors. 

Generally the difference between the temperatures or weights of the two water 
columns is enough to maintain a slow, but under favorable circumstances sufficiently 



Fig. 466.—Cooling System for Horizontal Engine, 
Natural Circulation. 


See Dingler’s Polyt. Journal, 1861, p. 2. 






















AUXILIARIES 


301 


rapid circulation between engine cylinder and tank. Where the pipe lines are long 

or have elbows, and generally in large installations this method of producing circu¬ 

lation is not sufficiently dependable, and it is better to cause circulation mechanically 
by means of a rotary pump. The periodic addition of fresh water to the system 
to replace that lost by vaporization is most easily made automatic by the use of 
a float-regulating valve. 

2. Extended Surface Radiators. These are connected by short pipe lines, top and 
bottom, to the jacket space, and give good service in small installations if they can 
be erected in an airy place. The internal width of these radiators is from 3 to 4", 
and the usual allowance is from 30 to 40 sq.ft, of superficial surface per B.H.P. per 
day of 10 hours. 

3. Cooling Towers. These use either natural draft or fan draft to cool the hot 

jacket water. Cooling towers with natural draft are used in large gas-engine installa¬ 

tions only when absolutely necessary, on account of high first cost and large space 
required. For every 1000 lbs. of water cooled, they require, when 12 to 14 layers 
of plank or crib work are used, from 5 to 10 sq.ft, of floor space. 

Fan coolers are quite widely used for portable oil engines. The hot water enters 

these coolers at the top through a spray nozzle, flows down over 10 to 15 partitions 

or baffle plates, around which air is blown from below, and gathers in a reservoir 
at the bottom, cooler by from 30 to 45°. From here it flows to the cylinder and 
back to the spray nozzle. Oil locomobiles powered with 8 to 12 H.P. use fan 

blowers having from 12 to 16" wheel diameter. The entire housing is made of 
sheet steel, and covering a space of approximately 5 sq.ft., has a height of from 

4 to 5 ft. Corrugated crib or plank wwk is the best filling to use, brush filling 
should be avoided. The power consumed by the circulating pump and the fan 
varies from 3 to 6% of the full engine power. 

4. Spray Nozzles and Cooling Ponds. Spray nozzles, when the water can be 

supplied to them under natural head only, find application in their simplest form 

also in small installations. But outside of this, they are used only in the large 

installations because the power consumed by the nozzles when a certain higher 
pressure is maintained on them is considerable. Cooling ponds become available only 
when large unused plots of ground can be had and when the soil is of such nature 
that costly masonry or concrete work is not necessary. As far as the cooling of 
jacket water is concerned, these ponds are to-day installed only where circumstances 
compel their use. The waste water of scrubbers and other cleaning apparatus is, 
on the other hand, often made again available by the use of cooling and settling 

ponds. 

Purity of Jacket Water is not less important than purity of feed water for boilers. 
Although the formation of scale occurs only in certain places of the jacket space 

which are highly heated, the throwing down of earthy impurities in the water forms 
mud or slush and other incrustations which seriously interfere with the cooling action. 
Water, about the quality of which there is any doubt, should never be directly 
supplied to the jacket, but should be first allowed to settle out all of the impurities 
possible in settling vessels or tanks. 

To prevent the freezing of the water in the jackets when the engine is out of 
operation, certain chemicals, such as alcohol, etc., are sometimes added to the water. 
According to the experience gained in the operation of motor cars, wood alcohol is 
-both cheap and well adapted to this purpose. The proportion used is 3 to 5%. 


302 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 


V. Piping 


Regarding the general assembly or lay-out of pipe lines, see the assembly draw¬ 
ings in Part III, especially Plates VIII, IX, XXV to XXVIII, and XXX. Of 
course the general rules regarding the design and construction of pipe lines apply 
also to this case (ample cross-sections, as few bends as possible, and never sudden 
change of direction, means of allowing for expansion, means for draining of water 
and for expelling entrapped air, protection of walls of buildings against possible 
vibrations of any parts of the line, etc.). Care should be taken especially to see 
that the exhaust pipe is placed as far as possible from air and gas pipes in order 
to prevent the radiation from the former from heating up the fresh charge, thus 
reducing its density, and consequently also the engine capacity. If circumstances 
do not permit of a satisfactory separation of exhaust and fresh gas pipes, the effect 
above mentioned should be counteracted by suitable cooling arrangements, such as 
covering the pipes. 

1. Air-pipe Line. If the volume of the fuel is small as compared with that of 
the air used, a condition that generally holds for oil engines, the internal diameter 
of the pipe may be determined from 



(4) 


in which D = cylinder diameter in feet, S = stroke in feet, v = allowable air velocity 
in the pipe in feet per second, and n = revolutions per minute. Depending upon the 
length of the line, v may be assumed from 30 to 60 ft. per sec. For the latter 
value, which holds for lines up to about 35 ft. long, eq. (4) becomes 



(5) 


Lines of abnormal length should be computed by means of any of the general 
formulae, taking into account pipe friction, pressure drop, etc. 

If the fuel gas constitutes a considerable part of the charge, the value of d 
may be made correspondingly smaller (see also p. 200). 

The intake end of the air-pipe line should be located in a spot cool, dry, and 
free from dust. The intake should be protected or screened in so that coarse dust, 
moisture or other undesirable additions can neither be drawn in nor reach the interior 
of the pipe in any other way. Regarding the muffling of the intake, see p. 296. 

2. Gas-pipe Line. The internal diameter of gas pipe for the connection between 
engine and pressure regulator or supply valve, for illuminating gas, is given in 
Table 28, p. 304, provided the length of the pipe is not more than 25 to 35 feet. 

The rest of the line from the gas bag to the main-supply pipe could be smaller, 
since in this pipe the gas, which is taken from the gas bag by the engine period¬ 
ically, flows in a steady stream. On the other hand, this part of the line is in most 
cases considerably longer than the other, and on account of the added frictional 








AUXILIARIES 


303 


resistance thus encountered, it is usual to make this pipe of somewhat larger diam¬ 
eter than the former. 

The pressure regulator is adjusted to give a pressure of from .4 to .6" water, 

measured at the gas bag. The pipe supplying the ignition flame should be branched 

off ahead of the gas bag, and should have or internal diameter. 

Power-gas lines should be computed on the basis of the quantity of gas, C h 

(see p. 200) that flows through the pipes at every suction stroke of the engine. For 
lengths not exceeding 35 to 45 feet, what has been said above regarding air-pipe 

lines applies equally to these producer-gas lines. For lines of considerable length 

the pressure drop in the pipe may be taken at from .07 to .12" water for every 
100 ft. of pipe. 

For producer gas (Dowson gas) the size of the gas pipe near the engine is 

made at least one half of the inlet-valve cross-section. In long lines this diameter 
is gradually attained. Too close figuring on the sizes of gas pipe for producer-gas 
installations is bad, because the occasional deposits of dust and tar after a time 

tend to reduce the normal cross-sections of the pipe. Special care should also be 

taken to make the connection between engine and gas-cleaning apparatus as short 

as possible and with the least number of bends or changes of direction. Cleaning 

openings of ample size and easily accessible should be placed wherever possible, and 
finally, the line should be provided, near the engine, with a purge opening and a 
try-cock. The lines leading from the ventilating or purge opening should be of 

sufficient size, not under 1 inch, and should always lead directly to the open air. 

Even in a small installation the much-used scheme of clearing or purging the pipe 
line or system through the try-cock from the poor gas made during the heating-up 
period leads to bad contamination of the air in the room. The connecting in of 

cleaning apparatus in the gas-pipe line proper is not good practice, see p. 290. 

3. Exhaust-pipe Line. The cross-section of the exhaust pipe from the engine to 

the muffler may be made from 1.1 to 1.3 times the free cross-section of the exhaust 

valve, depending upon the length of the connections. On the other side of the muffler 

the line may be somewhat smaller than the valve cross-section, provided it is straight 
and of ordinary length. This statement, however, does not apply to 2-cycle scaveng¬ 
ing engines, in which the exhaust pipe should always be made as large as possible. 

The discharge of exhaust gas through the chimneys of living rooms should not 
be permitted. In the erection of the exhaust line, attention should be paid to axial 
expansion, to the possibility of drawing off any water from the lowest points of the 
line, and to satisfactory isolation of the pipe from fuel gas and fresh-air lines as 

well as from inflammable material. The first division of the pipe connecting with 
the engine, and in installations above about 100 H.P., all of the exhaust pipe in the 
engine room should be water-cooled, for which purpose the water from the cylinder 
jackets may be used. Concerning the muffling of the exhaust, see p. 297. 

4. Cooling-water Lines. Table 28 gives the diameter of the water pipe to and 
from the cooling tank when natural circulation is used. When a special circulating 
pump is used, the pipe sizes may be reduced to almost those given for fresh-water 
cooling in the same table. The size of the inlet pipe for fresh-water cooling, i.e., 
where the hot water is wasted, may be determined from 

d = \/.023 N n in., .(6) 

wdien the water consumption is placed at the liberal figure of 10 gallons per B.H.P. 
hour, and the velocity in the pipe is assumed at 3 ft./sec. The size of the outlet 




304 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

pipe is made somewhat larger, from 1.25 to 1.75 d, depending upon length and drop. 
Table 28 also gives some average figures for this case. 

The branch lines leading to the other engine parts must be computed according 
to the quantity of water they must carry. The sum of the cross-section of all 
the branch pipes of course must not exceed the cross-section of the main pipe. Each 
branch line is furnished with its own regulating valve which is set once for all for 
the proper supply. When the engine is shut down, the water is shut off by a valve 
in the main line, while all of the valves in the branch line are left untouched. The 
water outlets should be free and so arranged near the engine that the hot-water 
temperatures can be easily read. Thorough draining of cylinder jacket and pipes 
is of importance on account of danger from frost and should be emphasized in the 
directions for operation. 


Table 28 

DIMENSIONS OF PIPE LINES FOR VARIOUS SIZES OF ENGINES 


Engine 

Capacity, 

B.H.P. 

Illuminating 
Gas Pipe. 

Size of Gas 
Meter, Number 
of Lights. 

Fresh Water 
Pipe, Inlet. 

Fresh Water 
Pipe, Outlet. 

Pipe to Cooling 
Tanks, Natural 
Circulation. 

Compressed 

Air Pipe Line, 
for Starting. 

2 

Inches. 

1 

20 

Inches. 

4 

Inches. 

Inches. 

14 

Inches. 

4 

1 

30 

4 

f 

14 


6 

1 

40 

+ 

4 

if 


8 

U 

50 

i 

4 

if 


10 

U 

60 

i 

4 

l 

2 


12 

H 

80 

f 

l 

24 


15 

H 

100 

1 

l 

24 


20 

2 

150 

f 

l 

2l 

14 

25 

2 

150 

4 

14 

2f 

14 

30 

24 

200 

4 

14 

2] 

14 

35 

24 

200 

l 

14 

3 

14 

40 

3 

250 

l 

14 

3 

14 

50 

3 

250 

H 

14 

34 

If 

60 

34 

300 

H 

14 

34 

If 

70 

34 

300 

14 

if 


2 

80 

34 

350 

14 

if 


2 

100 

4 

400 

if 

2 


24 

125 

4 

400 

if 

2 


24 

150 

44 

450 

2 

24 


24 


VI. Gas Meters, Gas Bags and Pressure Regulators 

These auxiliaries are used almost exclusively for illuminating-gas engines. Most 
gas-engine builders do not manufacture them themselves, but obtain them from special 
factories. For that reason a consideration of the details of construction is probably 
not necessary. 

1. Gas Meter. Gas meters are designated according to the number of lights 
they will supply. Table 29 gives the approximate capacity of each size of meter 
in cubic feet of gas per hour. The size of the meter shou-ld be based upon the 
-consumption at maximum engine capacity. 



















AUXILIARIES 


305 


Table 29 


SIZES OF GAS METERS 


3 lights, capacity = approximately 17.5 cu.ft. per hr. 


5 

< < 

( i __ 

( ( 

28.0 

10 

( ( 

( ( __ 

( C 

52.5 

20 

i < 

l ( __ 

t ( 

100.0 

30 

({ 

< < 

( ( 

140.0 

40 

( ( 

( ( 

( t 

195.0 

60 

( ( 

( ( __ 

( < 

280.0 

80 

11 

( ( 

i i 

385.0 

100 

( i 

i ( 

( ( 

490.0 

150 

( < 

( ( 

( ( 

700.0 

200 

( ( 

( i _ 

( ( 

1000.0 


< i 


The usual practice is for the gas company to furnish the meters and for the 
consumer to instal them. 

2. Gas-pressure Regulators. The purpose of these regulators is to maintain the 

gas pressure in the supply pipe at a certain level, irrespective of the variation in 

the main pipe. They are usually set to maintain a pressure of from .5" to .75" 

water, by suitably weighting the float. 

3. Gas Bags. These serve the purpose of equalizing the pressure fluctuations in 

the supply pipe due to the periodic aspiration of gas by the engine, and are there¬ 
fore connected in directly beyond the pressure regulator , i.e., between the latter and 
the engine. For the smaller engines, up to say 30 H.P., a single bag gives satis¬ 
factory service; in the larger installations several of them should be connected in 
series, or suction reservoirs of sufficient capacity should be connected in directly 

beyond the pressure regulator. 

The ends of the gas pipes should fill the rubber hose connections of the bag 

tightly, and should extend into the bag proper anywhere from 4" to 8". They 
should be strongly rounded off to prevent a cutting or wearing through of the rubber 
walls. The bag should not be called upon to support any part of either the inlet 
or outlet pipe. A good way is to use a continuous pipe passing in and out of the 
bag and to slot or perforate this pipe at the middle of the bag to give sufficient 
free passage for the gas. To protect the rubber against oil or possible external 
injury it is well to encase the bag in wood. 


Table 30 

APPROXIMATE DIMENSIONS OF RUBBER BAGS FOR REGULATING PRESSURE IN 

GAS MAINS 


Capacity of Engine. 

B. H. P. 

Width of Bag. 

Length of Bag. 

Length of Hose 
Connections. 

Internal Diameter of 
Hose Connections. 


Inches. 

Inches. 

Inches. 

Inches. 

.5 

13.0 

18.0 

3.25 

.80 

1-2 

16.0 

20.0 

3.75 

1.00 

3-5 

20.0 

24.0 

4.00 

1.20 

6-8 

22.0 

28.0 

4.50 

1.60 

10-12 

26.0 

30.0 

4.75 

1.80 

16-20 

28.0 

32.0 

4.75 

2.00 

25 

30.0 

34.0 

5.25 

2.40 

30 

32.0 

38.0 

5.25 

2.80 














306 DESIGN AND CONSTRUCTION OF INTERNALrCOMBUSTION MOTORS 


VII. General Machine Parts 

The following tables give data concerning dimensions of a number of simple 
general machine parts which, although in constant use in the construction of gas 
engines, are not usually manufactured by the engine builder but bought by him in 
the market. The dimensions of such parts have become generally standard and the 
gas-engine designer therefore must, in their use, conform to general practice. For 
such parts as have not become generally standard, average practical values will be given. 

1. Screws and Studs. Connections under heavy loads should only be made up by 
means of machine screws, because they are cheap and simple, and above all most 
reliable. Screws less than §" in diameter may be twisted off by careless setting up, 
hence their use in important places should be avoided. Studs are screwed into their 
seats so that the length of thread in position is at least 1.5 d when placed in machin¬ 
ery steel and from 1.8 to 2d when screwed into cast iron or brass. The hole, when not 
passing all the way through, should be at least f of the screw diameter deeper than this. 

Table 31 

U. S. OR SELLER’S SYSTEM OF SCREW-THREADS 


Bolts and Threads. 


Hexagonal Nuts and Heads. 


Long 


Diameter 

of 

Bolt. 

Threads 

per 

Inch. 

Diameter 
at Root of 
Thread. 

Width 

of 

Flat. 

Area of 
Bolt Body 
in 

Square 

Inches. 

Area at 
Root of 
Thread in 
Square 
Inches. 

Short 

Diam., 

Rough. 

Short 

Diam. 

Long 

Diam., 

Rough. 

Thick¬ 

ness, 

Rough. 

Thick¬ 

ness, 

Finish. 

Diameter 

Square 

Nuts, 

Rough. 

Inches. 

i 

20 

Inches. 

.185 

Inches. 

.0062 

.049 

.027 

Inches. 

4 

Inches. 

ft 

Inches. 

ft 

Inches. 

i 

Inches. 

ft 

Inches. 

7 

To 

ft 

18 

.240 

.0074 

.077 

.045 

ft 

ft 

n 

ft 

f 

1 0 

1 2 

f 

16 

.294 

.0078 

.110 

.068 

ft 

' f 

n 

f 

ft 

fi 

ft 

14 

.344 

.0089 

.150 

.093 

ft 

ft 

-ft 

ft 

f 

1ft 

f 

13 

.400 

.0096 

.196 

.126 

f 

If 

i 

f 

ft 

1ft 

ft 

12 

.454 

.0104 

.249 

.162 

§f 

ft 

H 

ft 

f 

1ft 

f 

11 

.507 

.0113 

.307 

.202 

ft 

l 


f 

ft 

If 

f 

10 

.620 

.0125 

.442 

.302 

H 



f 

if 

1ft 

1 

9 

.731 

.0138 

.601 

.420 

i ft 

if 

iff 

f 

ft 

2ft 

l 

8 

.837 

.0156 

.785 

.550 

if 

ift 

if 

1 

ft 

2ft 

if 

7 

.940 

.0178 

.994 

.694 

ift 

if 

2* 

H 

1ft 

2ft 

il 

7 

1.065 

.0178 

1.227 

.893 

2 

iff 

2A 

H 

1ft 

2ft 


6 

1.160 

.0208 

1.485 

1.057 

2ft 

2£ 

2ft 

If 

1ft 

3ft 

If 

6 

1.284 

.0208 

1.767 

1.295 

2f 

2ft 

2f 

If 

1ft 

3ft 

If 

5f 

1.389 

.0227 

2.074 

1.515 

2ft 

2b 

2ft 

If 

1ft 

3f 

If 

5 

1.491 

.0250 

2.405 

1.746 

2f 

2 ft 

3* 

G 

1ft 

3ft 

If 

5 

1.616 

.0250 

2.761 

2.051 

2ft 

2f 

3ft 

U 

1ft 

4ft 

2 

44 

1.712 

.0277 

3.142 

2.302 

3* 

3ft 

3» 

2 

1ft 

4ft 

2f 

4 

1.962 

.0277 

3.976 

3.023 

3f 

3ft 

4^ 

2f 

2ft 

4ft 

2f 

4 

2.176 

.0312 

4.909 

3.719 

3f 

3ff 

4f 

24 

2ft 

5ft 

2f 

4 

2.426 

.0312 

5.940 

4.620 

4* 

4ft 

4ft 

2f 

2ft 

6 

3 

3f 

2.629 

.0357 

7.069 

5.428 

4f 

4ft 

5f 

3 

2ft 

6f£ 

3f 

3f 

2.879 

.0357 

8.296 

6.510 

5 

4 if 

5M 

3f 

3ft 

7ft 

3f 

31 

3.100 

.0384 

9.621 

7.548 

5f 

5ft 

6^ 

3f 

3ft 

7ft 

3f 

3 

3.317 

.0413 

11.045 

8.641 

5f 

5ft 

6ff 

3f 

3ft 

8£ 

4 

3 

3.567 

.0413 

12.566 

9.993 

6£ 

6ft 

7A 

4 

3ft 

8ft 

41 

2f 

3.798 

.0435 

14.186 

11.329 

64 

6ft 

7A 

4f 

4ft 

9ft 

4f 

2f 

4.028 

.0454 

15.904 

12.743 

6f 

m 

7ft 

4f 

4ft 

9f 

4f 

2f 

4.256 

.0476 

17.721 

14.226 

7* 

7ft 

8ft 

4f 

4ft 

lOf 

5 

2£ 

4.480 

.0500 

19.635 

15.763 

7f 

7ft 

8ft 

5 

4ft 

10ft 

5f 

4 

4.730 

.0500 

21.648 

17.572 

8 

7ft 

9ft 

5f 

5ft 

lift 

5^ 

2f 

4.953 

.0526 

23.758 

19.267 

8f 

8ft 

9ft 

54 

5ft 

Hf 

5f 

2f 

5.203 

.0526 

25.967 

21.262 

8f 

8ft 

10ft 

5f 

5ft 

12f 

6 

21 

5.423 

.0555 

28.274 

23.098 

9£ 

9ft 

10ft 

6 

5ft 

12ft 























AUXILIARIES 


307 


Table 32 

ALLOWABLE LOAD IN POUNDS ON SCREWS OF VARIOUS SIZES, ACCORDING TO BACH. 1 


1 

2 

3 

4 

5 

6 

7 

8 

Dimensions. 

Allowable Load. 

External 

Diameter, 

Inches. 

Area at 
Bottom of 
Thread, 
Squarelnches 

Screws under Tensile Stress 
only. 

Screws under Tensile and 
Torsion Stress. 

Screws set up comparatively 
hard in the beginning and 
further drawn up during 
operation, as Flange Bolts 
for example. 

Lbs. 

Lbs. 

Lbs. 

Lbs. 

Lbs. 

Lbs. 

1 

4 

.027 

185 

230 

138 

175 

32 

40 

A 

.045 

310 

388 

233 

291 

74 

88 

1 

.068 

467 

583 

348 

435 

132 

165 

A 

.093 

640 

800 

480 

600 

216 

271 

4 

.126 

825 

1030 

617 

775 

324 

407 

1 

.202 

1385 

1730 

1040 

1300 

623 

782 

a 

4 

.302 

2070 

2590 

1560 

1940 

1070 

1305 

l 

.420 

2880 

3600 

2150 

2700 

1610 

2020 

1 

.550 

3780 

4730 

2830 

3530 

2110 

2650 

H 

.694 

4750 

5950 

3570 

4450 

2680 

3320 

H 

.893 

6100 

7640 

4580 

5730 

3430 

4280 

if 

1.057 

7200 

9060 

5420 

6790 

4070 

5090 

14 

1.295 

8870 

11050 

6650 

8330 

4980 

6230 

If 

1.515 

10000 

12550 

7550 

9400 

5650 

7060 

If 

1.746 

11900 

14900 

8950 

11700 

6740 

8400 

H 

2.051 

13550 

16920 

10100 

13300 

7620 

9550 

2 

2.302 

15750 

19650 

11800 

15450 

8850 

11100 

2} 

3.023 

20000 

25000 

14950 

19550 

11200 

14050 

2i 

3.719 

25400 

31900 

19000 

24900 

14300 

17900 

21 

4.620 

30400 

38100 

22800 

29800 

17100 

21400 

3 

5.428 

37100 

46500 

27800 

36500 

20900 

26000 


1 The original table was based on Whitworth thread, but the figures have been directly trans¬ 
posed, since the differences between the bottom areas of this and the Seller’s thread are not great. 


Note. Columns 3, 5, and 7 apply to screws of steel of average quality; columns 4, 6, and 8 
to screws of superior grade. The stress per square inch of area at bottom of thread has been 
taken as follows: 


Column. 3 4 5 6 7 8 

6800 8500 5100 6400 3820 4650 lbs. per sq.in., 

except for screws from \ in. to f in., for which the stress has in each case been taken 10% less. 
Where first-class soft steel is employed the values in columns 7 and 8 may be exceeded 30%. 

For gas thread, see Table 34, p. 309. 



































308 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 

• > 

2. Keys. The material used is soft or crucible steel. Keys are tapered along 
the top surface only and the taper amounts to from T V to When two sunk, 

flat, or saddle keys are employed they are placed at an angle of 120°, unless pecu¬ 
liarities of hub construction, etc., compel the use of smaller angles. 


Table 33 

DIMENSIONS OF KEYS 



Figs. 467a-467d. 





Flat on Saddle Keys. 

Sunk Keys. 


Diameter 

of 



Keys. 

Key Seat 

Depth in 


Shaft. 

Width, b. 

Thickness, a. 







Width, b. 

Thickness, a. 

Shaft. 

Bore. 




Inches. 

Inches. 

Inches. 

Inches. 

Inches. 

Inches. 

Inches. 


.75 

.40 

.20 

.40 

.25 

.10 

.15 


1.00 

.45 

.20 

.45 

.25 

.10 

.15 


1.25 

.50 

.20 

.50 

.30 

.10 

.20 

r ET 

1.50 

.55 

.25 

.55 

.35 

.10 

.25 

c CN1 

1.75 

.65 

.25 

65 

.40 

.125 

.275 

•Bid 

2.00 

.70 

.25 

.70 

.40 

.125 

.275 

a <N 

2.25 

.80 

.25 

.80 

.45 

.15 

.30 


2.50 

.85 

.30 

.85 

.50 

.15 

.35 

M 

2.75 

.90 

.30 

.90 

.50 

.15 

.35 

HfN fc- rr 

l -g * 

^ 5 ^ 

3.00 

.95 

.35 

.95 

.55 

.15 

.40 

3.50 

1.10 

.35 

1.10 

.60 

.20 

.40 

“ o S 

4.00 

1.20 

.50 

1.20 

.70 

.20 

.50 

^ o 

4.50 

5.00 

Suitable for light shafts only. 

1.35 

1.50 

.75 

.80 

.20 

.25 

.55 

= CM 



.55 

•d Cu 

6.00 

Double Keys. 

1.70 

.95 

.30 

.65 

w CO 

7.00 



2.00 

1.10 

.35 

.75 


8.00 

1.00 

.80 

2.25 

1.20 

.40 

.80 

o jr • 
o £ bjo 

9.00 

1.15 

.90 

2.50 

1.45 

.45 

1.00 

£ a rT 

x 

10.00 

1.30 

1.00 

2.75 

1.65 

.50 

1.15 

* v cT 

12.00 

1.55 

1.25 

3.20 

2.00 

.55 

1.45 

u. o 

o o M 

14.00 

1.75 

1.40 

3.70 

2.25 

.70 

1.55 


16.00 

1.95 

1.50 

4.10 

2.45 

.80 

1.65 

a 

o 

18.00 

2.10 

1.60 

4.50 

2.60 

.90 

1.70 

r- & C* 

20.00 

2.20 

1.70 

4.80 

2.80 

1.00 

1.80 




































AUXILIARIES 


309 


3 . Gas Pipe, Pipe Fittings, and Flanges. 


Table 34—DIMENSIONS OF WROUGHT-IRON WELDED PIPE—U. S. Standard 


Nominal 

Inside 

Diameter. 

Actual 

Outside 

Diameter. 

Actual 

Inside 

Diameter. 

Thickness 

of 

Metal. 

Internal Area. 

External Area. 

Weight of 
Pipe per 
Lineal Foot. 

Number of 
Threads 
per Inch. 

Inches. 

Inches. 

Inches. 

Inches. 

Square Inches. 

Square Inches. 

Lbs. 

No. 

i 

.405 

.270 

.068 

.057 

.1288 

.24 

27 

i 

.540 

.364 

.085 

.104 

.2290 

.42 

18 

% 

.675 

.493 

.091 

.191 

.3578 

.56 

18 

i 

.840 

.622 

.109 

.304 

.554 

.84 

14 

1 

4 

1 .050 

824 

.113 

.533 

.866 

1.12 

14 

1 

1.315 

1.048 

.134 

.861 

1.358 

1.67 

111 

H 

1.660 

1.380 

.140 

1.496 

2.164 

2.24 

111 


1.900 

1.610 

.145 

2.036 

2.835 

2.68 

1 H 

2 

2.375 

2.067 

.154 

3.356 

4.430 

3.61 

ni 

24 

2.875 

2.468 

.204 

4.780 

6.492 

5.74 

8 

3 

3.500 

3.067 

.217 

7.383 

9.621 

7.54 

8 

34 

4.000 

3.548 

.226 

9.887 

12.566 

9.00 

8 

4 

4.500 

4.026 

.237 

12.730 

15.904 

10.66 

8 

44 

5.000 

4.508 

.246 

15.961 

19.635 

12.34 

8 

5 

5.563 

5.048 

.259 

19.986 

24.301 

14.50 

8 

6 

6.625 

6.065 

.280 

28.890 

34.472 

18.76 

8 

7 

7.625 

7.023 

.301 

38.738 

45.664 

23.27 

8 

8 

8.625 

7.981 

.322 

50.027 

58.426 

28.18 

8 

9 

9.625 

8.937 

.344 

62.730 

72.760 

33.70 

8 

10 

10.75 

10.018 

.366 

78.823 

90.763 

40.06 

8 

11 

11.75 

11.000 

.375 

95.033 

108.434 

45.02 

8 

12 

12.75 

12.000 

.375 

113.098 

127.677 

49.00 

8 

13 

14 

13.25 

.375 

137.887 

153.938 

54.00 

8 

14 

15 

14.25 

.375 

159.485 

176.715 

58.00 

8 

1 15 

16 

15.25 

.375 

182.665 

201.062 

62.00 

8 


18 

17.25 

.375 

239.706 

254.470 

70.00 

. , 


20 

19.25 

.375 

291.040 

314.159 

78.00 



22 

21.25 

.375 

354.657 

380.134 

85.00 

# , 


24 

23.25 

.375 

424.558 

452.390 

93.00 




Table 35 

DIMENSIONS OF BRASS UNIONS FOR HIGH-PRESSURE 
COPPER PIPE 

All Dimensions in Inches 

Material of the union itself is brass or bronze, that of the cone is copper 


Dian 
of r 

In ¬ 

ter¬ 

nal . 

leter 

ipe. 

Ex¬ 

ter¬ 

nal . 



Small 

Diameter 

of 

Hexagon. 

















a 

a' 

b 

b' 

c 

c' 

d 

e 

/* 

g 

h 

i 

k 

l 

m 

n 

O 

p 

9 

r 

S 

t 

.16 

.28 

.44 

.48 

.92 

.76 

.68 

.32 

r 

1.44 

1.80 

1.20 

.60 

.28 

1.00 

.60 

With seat angle 

1.60 

.60 

1.72 

.28 

.44 

.60 

.64 

1.08 

.92 

.84 

.40 

i" 

1.60 

1.80 

1.20 

.60 

.36 

1.00 

.60 

continues 

1.60 

.60 

1.80 

.40 

.56 

.72 

.76 

1.32 

1.08 

1.00 

.48 

1" 

1.76 

1.80 

1.20 

.60 

.40 

1.00 

.60 

straight to the 

1.60 

.60 

1.84 

.48 

.64 

.84 

.88 

1.32 

1.32 

1.16 

.56 

H" 

1.92 

1.80 

1.20 

.60 

.48 

1.00 

.60 

tube up to 

1.60 

.80 

2.12 

.56 

.76 

.96 

l .00 

1.44 

1.32 

1 .32 

.64 

H " 

2.08 

1.80 

1.20 

.60 

.52 

1.00 

.60 

a — 

5b 


1.60 

.80 

2.10 

.64 

.84 

1 .04 

1.08 

1 .60 

1.44 

1 .40 

.68 

H" 

2.08 

1.80 

1.20 

.60 

.52 

1.00 

.60 

.20 

.80 

.60 1.60 

.80 

2.16 

.80 

1.00 

1.20 

1.24 

1 .80 

1.60 

1.60 

.76 

ll" 

2.32 

1.80 

1.20 

.60 

.52 

1.00 

.80 

.20 

1.00 

.60 1.80 

1.00 

2.16 

1.00 

1.20 

1.44 

1.48 

2.00 

1.80 

1.88 

.88 

11 " 

2.64 

1.88 

1.20 

.60 

.56 

1.00 

1.00 

.20 

1.20 

. 60 ; 2.00 

1.20 

2.20 

1.20 

1.40 

1.68 

1.72 

2.16 

2.00 

2.16 

1.00 

2 

2.88 

1.88 

1.20 

.60 

.60 

1.00 

1.20 

.20 

1.40 

.60 2.20 

1.40 

2.24 


* Pipe Thread 

























































































Table 36 

DIMENSIONS AND WEIGHT OF STANDARD PIPE FITTINGS 
Compiled mostly from data given by Crane Co. 


310 DESIGN AND CONSTRUCTION OF INTERNAL-COMBUSTION MOTORS 
























































































AUXILIARIES 


311 



Table 37 

DIMENSIONS OF STANDARD PIPE FLANGES 
(Millar & Son) 

All Dimensions in Inches 


Size of Pipe, 
Inches. 

External 

Diameter. 

A 

Thickness. 

B 

Diameter 
of Hub. 

C 

Length of 
Thread. 

D 

Diameter 

of 

Bolt Circle. 

Number 

of 

Bolts. 

Size 

of 

Bolts. 

Length of 
Bolts for 
Standard. 

1 

4 

16 

1M 

\\ 

3 

4 

A 

11 

H 

4* 

1 

2A 

a 

4 

3| 

4 

A 

11 

H 

5 

A 

2f 

i 

31 

4 

1 

U 

2 

6 

t 

3* 

i 

4| 

4 

t 

2 

2\ 

7 

H 

3f 


51 

4 

t 

21 

3 

71 

a 

4 

4A 

u 

6 

4 

t 

21 

3* 

Si 

H 

41 

i A 

7 

4 

t 

24- 

4 

9 

it 

5f 

1A 

71 

4 


21 

41 

91 

it 

5A 

H 

7! 

8 

1 

3 

5 

10 

it 

6A 

ift 

Si 

8 

i 

3 

6 

11 

l 

7A 

i* 

91 

8 


3 

7 

121 

i* 

8f 

H 

101 

8 

i 

31 

8 

131 

H 

9H 

H 

lit 

8 

I 

31 

9 

15 

H 

10| 

if 

13i 

12 

i 

34 

10 

16 

l A 

nit 

U 

141 

12 

1 

31 


Table 38 

DIMENSIONS SUITABLE FOR OVAL FLANGES 
All Dimensions in Inches 


Size of pipe . 

1 

4 

t 

1 


a 

t 

1 

11 

H 

1 2 

1 4 

2 

21 

3 

Width of flange, a . 

1.40 

1.60 

1.92 

2.08 

2.32 

2.64 

3.04 

3.36 

3.76 

4.10 

4.80 

5.56 

Length of flange, o' . 

2.64 

2.80 

3.44 

3.60 

3.76 

4.00 

4.64 

4.96 

5.46 

6.16 

6.90 

7.76 

Center to center of holes, b . 

1.60 

1.68 

2.16 

2.32 

2.48 

2.72 

3.20 

3.44 

3.76 

4.24 

4.87 

5.60 

Diameter of bolt hole, c . 

A 

A 

1 

1 

1 

1 

t 

a. 

8 

5 

8 

13. 

16 

it 

it 

Thickness of flange, e* . 

.36 

.36 

.36 

.40 

.40 

.40 

.40 

.40 

.50 

.50 

.60 

.72 

Radius, /1 . 

2.32 

2.16 

2.68 

2.52 

2.42 

2.36 

2.72 

2.80 

3.08 

3.60 

3.84 

4.24 

Radius, g . 

.52 

.56 

.64 

.64 

.64 

.64 

.72 

.76 

.84 

.96 

1 .00 

1.08 


* Applies to steel flanges, for cast iron flanges, thickness = appr. l§e. 

a 

t Applies to Type I, for Type II mean radius = -q. 



























































312 DESIGN AND CONSTRUCTION OF INTERNAUCOMBUSTION MOTORS 


4. Helical Springs. In the following table, which was printed in Power, July 
1907, P represents the maximum safe load for helical springs made of round-steel 
wire, D is the pitch diameter of the spring in inches to the center of the wire, and 
/ the corresponding deflection of the spring (tension or compression) in inches for 
one coil. 

The table is based upon the following general equations: 



K d = . 392 


d 3 

D 


Kd lbs., 


. 64r 3 P. 8 D 3 P. . 

d*E l * ~ d*E im ' 


d = diameter of the wire in inches; 

E = torsional modulus, assumed at 12600000; 
i = number of windings = 1.0 in the table; 

Kd = safe fiber stress, has been assumed to vary for the different sizes of wire 
as shown in the table. 1 


Regarding the determination of the principal dimensions of valve springs, see 
p. 203; concerning the law for the variation of load with deflection, see the force 
diagram, Fig. 257. 


1 The computer of this table states that it has been adopted by a number of designers and 
found useful. It seems to the translator, however, that the maximum fiber stress assumed is rather 
high for all sizes of wire from to H". In his own practice he would use about 60 000 lbs. for the 
H" wire, and increase it by steps up to 100 000 lbs. for the wire. If it is decided to use any 
other stress than that assumed in the table, the corrections for P and / are very quickly made, for 
since Kd occurs in the first power in the equation for P, and P occurs in the first power in the 
equation for /, it is only necessary to multiply the figures in the table by the ratio of the stresses, 
taken in the proper sense. For instance 

From the table, for a stress assumed at 90 000 lbs., 5 = 1.0, and pitch diameter = 6.0", 
P = 5870, / = .805". 

If the stress had been assumed at 60 000 lbs., then 


60 000 
90 000 


5870 = 3900 


lbs., 


and 


60 000 
90 000 


.805 = .536". 







Table 

SAFE LOADS AND DEFLECTIONS OF Hftu'at cm,. 

Pitch Diameter of Spkinus=Z>. " j ® I EEL SPRINGS 


JL" 

16 


A' 


6.25 

4.21 

3.21 

. 00606 

.01360 

.02425 

9.50 

6.30 

1 .75 

.00530 

.01602 

.02130 

12.28 

8.15 

6.08 

.00487 

.01085 

.01950 

17.30 

11.50 

S.63 

00435 

.00925 

.01740 

.031 

15.63 

12.70 

.0088 

.01565 

.032 

20.80 
.0103 

15.51 

.01735 

| 20 

.035 

20.10 

.0168 

19 

.041 

32.60 

.01425 


v 

8 


IS 

17 


.047 

.054 


A' 


150000 


1 V 

16 

15 


2.50 
.03765 
3.71 
.03610 
4.90 
.03050 
6.85 
.02700 
9.350 
02445 
12 To 

.02850 
16.20 
.0262 
26.20 
.02225 
39.12 
.01945 
59.40 
.01685 


.062 

.063 

.072 


14 


13 

h 


JL" 

1 " 

TO 

4 

2.06 

1.545 

.05433 

.09700 

3.15 

2.38 

.04765 

.08480 

4.025 

3.05 

.04375 

.07800 

5.76 

4.32 

. 03900 

.06950 

7.820 

5.870 

.0352 

.06227 

10.32 

7.700 

.0410 

.0766 

13.45 

10.08 

.0377 

.0671 

21.65 

16.20 

.0320 

.0570 

32,60 

24.55 

.0280 

. 049S 

49.60 

37.25 

.0244 

.0435 

74.90 

56.10 

.0212 

. 0378 

78.24 

58.70 

.0208 

.0370 

117.3 

80.75 

.0183 

.0326 

.080 

121.0 


.0294 

092 

171.0 


.02365 

.093 

178.0 


.02342 

1 12 

.105 

Jil 

.120 


JL" 

16 


140000 


i 

S’ 

10 


1.24 
. 15300 
1.845 
.13250 
2.50 
.12200 
3 . 46 
.14300 
4.670 
.0980 
6.140 
. 1150 
8.100 
. 1055 
13 04 
.0890 

19.60 
.0782 
29.70 
. 0678 

44.90 
.0587 

46.90 
.0578 
70.05 
.0509 

96.60 
. 0455 

1367.T 

.0.371 
142.0 
.03651 
204.0 
.0328 
303.5 
. 0284 

.125 

.135 


:s it 

s 


A 


1.042 
.21800 
1.521 
. 19100 
2.065 
. 17500 
2.83 
. 15300 
3.910 
. 1410 
5 085 
. 1658 
6.660 
. 1490 
10.85 
.1280 
16.445 
.1125 
24.66 
.0980 
37.30 
.0849 
39.20 
.0835 
58.70 
.0734 
80.50 
.0652 
113.5 
.0534 
118.2 
.0528 
170.0 
.0472 
253.0 
.0411 
286.0 
.0394 
359.0 
.036 4 

.148 

.156 


3 

10 

6 


X" 

16 

. 895 
. 29700 
1 .325 
.26000 
1.74 
.23900 
2.39 
.21350 
3.360 
.1920 
4.440 
.2245 
5 750 
.2008 
9.270 
. 1750 
13.95 
. 1530 
21.20 
. 1334 
32.00 
. 1158 
33.90 
.1135 
50.20 
.1000 
69.15 
.0900 
97.60 
.0728 
99.50 
.0715 
146.0 
.0640 
217.5 
.0559 
245.0 
.0532 
309.0 
.0495 
408.0 
.0450 
480.0 
.0429 

.162 

.177 


. 1 " 


Load in pounds = P 
Compression per coil in inches* 


.187 

.192 


125000 


£ 

4 

3 


1.52 



.31250 



2.18 

5" 


. 27800 

S 


2.935 

2.331 

) 

.2510 

.3801 

1 

3.780 

3.131 

1 3" 

. 2940 

.4601 

* 

4.960 

4.051 

3.390 

.2760 

.4201 

.6080 

8.100 

6.521 

5.350 

. 2290 

.3621 

.5125 

12.30 

9.801 

8.100 

. 2000 

■ 3ii5 

.4491) 

IS. 55 

14.75 

12.40 

. 1740 

.2730 

.3900 

28.00 

22.40 

18.65 

.1510 

. 2350 

.3400 

29.40 

23.50 

19.60 

. 1480 

.2330 

.3350 

43.60 

35.20 

29.00 

.1304 

.2038 

.2940 

60.40 

48.28 

40.10 

.1173 

.1835 

2625 

85.50 

68.90 

57.30 

.0950 

.1488 

.2140 

89.00 

71.20 

69.10 . 

.0938 

. 1463 

.2110 

127.5 

102.0 

S5.40 

.0838 

. 1225 

.1880 

190.0 

152.5 

126.8 

.0732 

.1140 

.1740 

214.0 

171.5 

143.0 

.0696 

. 1090 

. 1690 

270.0 

217.1) 

171.0 

.0645 

. 1060 

.1458 . 

356.0 

285.0 

237.5 2 

.0503 

.0921 

. 1325 

418.0 

333.0 

270.02 

.0560 

.0875 

.1257 . 

468.0 

376.0 

311.5 2 

.0540 

.0842 

.1217 . 

608.0 

487.0 

106.0 3 

.0497 

.0772 

.1150 . 

642.0 

522.0 

426.03 

.0415 

.0650 

.0935 

696.0 

556.0 

465.0 3 

.0408 

.0637 

.0917 . 


394.0 

579.0 4 


0590 

0850 . 


3" 

8 


4.570 

.6970 

6.920 

.6100 

10.58 

.5320 

16.10 

.4600 

16.80 

4550 

!5.00 

.40501 


1 


ir 


.218 


.244 


32 


812 

.0557 

895 

.0540 

1120 

.0499 

.250 

.263 

.281 


1 

O' 

16 


678.0 

.0802 

746.0 

.0780 

950 

.0718 

1027 

.0701 

1195 

.0665 

1450 

.0621 

.283 

.307 

.312 


Ji 

3 

0 


115000 


. 3590 
18.80 
2910 
>0.90 
2865 
’3.00 
2560 


213. : 


1980 


1 1 15 

76.5 


1512 


127: 


1154 
580 
. 1090 
640 
1062 
811 
.0980 
880 
.0957 
1125 
.0895 
1240 
.0850 
1264 
.0843 
1630 
.0780 
1575 
. 0703 

.331 

.343 

.362 


.6.140 
.8000 
9.255 
. 9650 
13.95 
. 6050 
14.70 
.5910 
21.90 
.5216 
30.10 
.4 700 
42.60 
.3800 
44.30 
.3760 
63.40 
. 3360 
95.20 
. 2960 
107.3 
. 2780 
135 
.2600 
178 
.2360 
208 
.2235 
234 
.2160 
305 
. 1980 
320 
.1660 
348 
.1634 
43 
.1514 
509 
.142 
560 
.1385 
711 
.1280 
760 
.1250 
895 
.1184 
1087 
.1110 
1110 
.1100 
1420 
. 1020 
1376 
.0920 
.1636| 
.0864 
1820 
.0825 
2140 
.0791 


ir 


8.230 

.8780 

12.50 
.7600 
13.20 
.7480 

19.50 
. 6520 
26.75 
.5935 

37.80 

4810 

39.60 
.4730 

56.60 
4250 

84.20 
.3685 
95.00 
.3530 
120.0 
.3275 
158.0 
. 2930 
185.3 
2820 
207.0 
.2730 


ir 


11.22 

.9470 

11.90 

.9300 

17.50 
8020 

24.20 
7350 

34.50 
5960 

35.70 

.5850 

51.10 

.5125 

76.20 
.4490 

85.20 
.4370 
108.5 

3990 

142.4 
.3700 
167.0 
.3500 

187.5 
.3380 


ir 

10.75 

1.122 

16.05 

.9860 

22.10 

.8860 


IV 

20420 

1.058 


17 


270 

243.4 

2510 

.3110 

288 

256 

2100 

.2600 

309 

278 


13 

32 


JL 

16 


.2055 
385.0 
.1910 
452 
.1805 
498 
.1750 
632 
.1610 
685 
.1585 
795 
.1493 
969 
. 1400 
985 
.1391 
1260 
.1290 
1220 
.1163 
1455 
.1090 
1620 
.1058 
1910 
.1000 
2110 
.0966 
2430 
.0920 


.406 

.430 

.437 


A 

0 


.2520 
346 
.2363 
408 
.2230 
44 
2130 
579 
2000 
617 
1915 
717 
1835 
869 
.1720 
886 
.1691 
1135 
.1565 
1100 
.1440 
1310 
.1356 
1452 
.1270 
1714 
.1238 
1940 
.1175 
2180 
.1140 
2400 
.1080 
2875 
.1041 
3000 
. 1002 


.460 

.468 


.490 


110000 


233 


254 


369 
. 269( 
407 
. 262( 
527 
. 240( 
560 
.2361 
652 
. 2241 
794 
209C 
805~ 
.207C 
1035 
.1920 
1000 
. 174C 
1187 
.1630 
1325 
.1560 
1560 
.1492 
1730 
.1442 
1984 
.1378 
2170 
.1340 
2400 
.1260 
2730 
.1243 


3065 

.1128 

3225 

.1111 

3675 


.500, 


20 28.6( 

J 11" 





X) .854 







10 29.6( 

27.34 






'0 .841) 

.9864 






10 42.6< 

38.94 

1 3" 




10 .755( 

> .8800 






10 63.5( 

58.50 

54.301 1 L" 



0 .657( 

1 .7680 

8930 0 




0 71.51 

65.80 

60.80:57.20 

2" 



0 .6261 

.7310 

,8550! .9810 




0 90.2( 

82.70 

77.20 71.80 

67.51 



8 ,581( 

.6800 

.7910 

.9080 

1.361 



0 118.5 

109.5 

102.0 

95.0 

86.0 

21" 


0 .530C 

.6200 

.7230 

.830 

.9450 



2 139 

128 

119.4 

111.6 

104.6 

92.70 


0 .5090 

.5880 

.6850 

.7850 

.8960 

.1121 


S 156 

143.6 

134 

125.5 

117 

103.6 

2\" 

0 .4880 

.5660 

.6630 

.7570 

.8631 

1.095 


5 205 

187.6 

174.2ll63.4 

152.5 

135.5 

122.0 

0 .4470 

.5220 

.6060 

1 .6950 

.7930 

1.001 

1.240 

2130 

197 

183 

170.5 

160 

142.5 

128.3 

3 .3730 

.4370 

.5100 

.5840 

.6650 

.8320 

1.041 

232.0 

214 

199 

186 

174 

154.5 

139.3 

) .3660 

.4280 

.4990 

.5710 

.6520 

.8250 

1.021 

2 288 

266 

247 

232 

216 

192.5 

173 

) .3420 

.3960 

.4620 

.5700 

.6070 

.7570 

.9430 

339 

310 

291 

270 

255 

225 

204 

.3210 

.3740 

.4370 

.4880 

.5700 

.7100 

.8910 

372 

345 

320 

299 

280 

248 

224 

) .3120 

.3720 

.4250 

.4860 

.6910 

.8660 

1.050 

475 

438 

406 

381 

356 

316 

284 

.2875 

.3360 

.3910 

.4490 

.5110 

.6460 

.8000 

513 

476 

440 

410 

385 

342 

308 

.2810 

.3280 

.3S50 

.4390 

.5000 

.6240 

.780 

598 

551 

501 

478 

448 

400 

359 

.2660 

.3120 

.3630 

.4160 

.4750 

.5920 

.7400 

724 

665 

620 

580 

543 

482 

435 

.2500 

.2920 

.3400 

.3900 

.4430 

.5620 

.6920 

740 

682 

634 

592 


492 

439 

.2460 

.2890 

.3380 

.3860 

.4-100 

.5590 

.6900 

945 

872 

810 

758 

710 

630 

568 

.2370 

.2670 

.3125 

.3560 

.40S0 

.5160 

.6380 

915 

845 

775 

733 

687 

610 

550 

.2070 

.2420 

.2830 

.3220 

.3680 

.4670 

.5770 

1090 

1000 

932 

870 

818 

725 

653 

. 1940 

. 2270 

.2640 

.3030 

.3460 

.4370 

.5410 

1214 

1120 

1040 

970 

910 

808 

728 

. 1860 

.2180 

.2560 

.2930 

.3300 

.4130 

.5220 

1430 

1318 

1220 

1142 

1070 

950 

858 

. 1778 

.2070 

.2430 

.2780 

.3170 

.4000 

.4950 

1580 

1458 

1354 

1265 

1185 

1058 

950 

. 1720 

.2010 

.2340 

.2680 

.3080 

.3820 

.4780 

1820 

1680 

1560 

1458 

1365 

1212 

1092 

. 1640 

.1915 

.2230 

.2560 

.2920 

.3690 

.4570 

2000 

1840 

1710 

1600 

1500 

1330 

1200 

.1595 

.1865 

.2170 

.2480 

.2840 

.3530 

.4440 

2610 

2210 

2050 

1918 

1798 

1598 

1440 

.1500 

.1752 

.2040 

.2340 

.2670 

.3380 

.4180 

2500 

2310 

2140 

2000 

1880 

1665 

1500 

.1480 

.1734 

.2010 

.2310 

.2640 

.3270 

.4120 

2800 

2580 

2400 

2230 

2100 

1865 

1680 

.1340 

.1570 

. 1830 

.2090 

.2390 

.3030 

.3740 

2940 

2725 

2530 

2375 

2210 

1970 

1770 

.1320 

.1548 

.1820 

.2065 

.2350 

.2950 

.3680 

3370 

3115 

2890 

2710 

2535 

2245 

2045 

. 1261 

.1480 

.1720 

.1965 

.2250 

.2840 

.3515 

| 3610 

3320 

3090 

2890 

2710 

2410 

2160 

.1236 

.1446 

.1685 

.1925 

.2200 

.2740 

.3470 

» 

4700 

4390 

4090 

3830 

3420 

3080 

10 

.1282 

.1493 

.1750 

.1950 

.2480 

.3060 


I 6100 

1 .1347 


5600 

.1540 


16 


5260 

.1760 

63 T 5 

. 1455 


100000 


4660 

.2180 

5650 

.1835 

7400 

.1735 


03" 

-¥ 

12673 
1 232 
158.0 
1.410 
185 
1.081 
203 
1.247 
259 
.9650 
281 
.9460 
326 
.8960 
395 
.8400 


4210 
.2750 
5090 
.2280 
6640 
.2150 
8420 
. 1920 
10530 
.1785 


402 

.8330 

516 

.7610 

500 

.6970 

594 

6530 

661 

6250 

778 

5980 

860 

5790 

990 

5500 

1090 

.5250 

1308 

.5030 

1365 

.4900 

1530 

.4470 

1610 

.4440 

1840 

.4240 

1970 

.4150 

2790 

.3720 

3825 

.3280 


144.6 

1.360 

169.5 
1.282 

187 

1.247 

237.5 
1.147 

266 

1.125 

298 

1.066 

362 

1.000 


3 1 


1 " 


222 
1.530 
256 
1.44, r 
310 
1.360 


370 

.9900 

473 

.9180 

460 

.8290 

545 

.7708 

608 

.7450 

714 

7130 

790 

6880 

910 

.6570 

1000 

.6400 

1200 

.6000 

1250 

.5930 


317 
1.350 
406 
1.247 
392 
1.128 
468 
.1056 
520 
1.025 
612 
.965 
678 
.932 
780 
. 890 
855 
.8670 
1028 
.8150 
1074 
.8100 


355 

1.632 

343 

1.475 

410 

1.302 

454 

1.320 

535 

1.264 

592 

1.225 

682 

1.165 

750 

1.136 

900 

1.068 

940 

1.056 


4V 


52S 

1.548 

607 

1.476 

666 

1.436 


4630 

.2745 

6030 

.2520 

7660 

.2320 

9550 

.2 150 

10600 

.1790 


90000 


1400 

.5360 

1472 

.5300 

1690 

.5060 

1810 

.4950 

2565 

.4400 

3505 

.5975 


4250 

.3275 

5540 

.2995 

7000 

.2760 

8700 

.2570 

9700 

.2170 

11780 

.2060 


1200 

.7290 

1265 

.7200 

1445 

.6880 

1550 

.6720 

2190 

.5950 

3000 

.5380 


H 

U 


3625 
.4340 
4745 
.4025 
6010 
.3760 
7500 
.3480 
8400 
.2900 
11000 
. 2760 
14400 
.2440 
19700 
.2200 


1058 

.9565 

1110 

.9430 

1268 

.9000 

1352 

.8800 

1913 

.7820 

2630 

.7050 


3195 

.5800 

4150 

.5325 

5260 

.4900 

6560 

.4560 

7160 

.3830 

8800 


SO 000 


1 a 

1 8 




800 

5" 



1.350 




835 

750 



1.337 

1.640 



932 

840 



1.210 

1.492 



985 

885 

5V' 


1.192 

1.475 



1125 

1015 

920 


1.139 

1.405 

1.700 


1205 

1082 

985 

6" 

1.115 

1.375 

1.650 


1710 

1535 

1395 

1280 

.9900 

1.220 

1.473 

1.754 

2340 

2110 

1913 

1750 

.8750 

1.058 

1.330 

1.583 

2825 

2560 

2330 

2125 

.7330 

.9080 

1.005 

1.305 

3690 

3325 

3025 

2770 

.6710 

.8320 

1.006 

1.19S 

4675 

4200 

3825 

3500 

.6200 

.7660 

.8800 

1.105 

7540 

5250 

4770 

4370 

.5770 

.7120 

.8600 

1.023 

6470 

5810" 

5290 

4850 

.4800 

.5910 

.7150 

.8550 

8600 

7050 

6330 

5870 

.4540 

.5610 

.6800 

.8050 

11230 

40100 

9200 

8400 

.4050 

.4960 

.6000 

.7180 

1530C 

13780 

12540 

11500 

. 364C 

.4420 

.5450 

.6480 

) 18101 

1615C 

14850 

13600 

) .2941 

.3641 

.4400 

.5220 

‘ 23501 

21400 

19300 

17700 

.2701 

.3320 

.402C 

.4800 


* Table from Power, July, 1907, 












































































































































































































































































PART III 


CONSTRUCTION, ERECTION AND TESTS OF MODERN 

COMBUSTION ENGINES 


INTERN AL- 


The following pages contaiii some detailed information on a selected number of 
the most recent types of internal-combustion engines, including in most cases results 
of the latest tests, together with lists of sizes and weights. Although the construc¬ 
tions discussed cannot perhaps in all cases be regarded as typical without qualification, 
taking them altogether they represent a very gratifying picture of the high develop¬ 
ment the gas-engine industry has to-day attained. 

The drawings, clear in execution and true to detail, hardly call for any extended 
descriptions. It is sufficient to point out the main details of the design, from which 
the manner of operation will in general be easily understood. In view of the fact 
that many of these types are used for oil as well as for gas fuel, the engines described 
in the following pages are not classified in any way according to fuel used, but the 
subject-matter is arranged simply according to makers. The sequence, has been 
decided, as far as possible, only with reference to the length of time the various 
engines have been on the market, there being absolutely no intention to establish 
an order of merit. 1 


A. STATIONARY ENGINES 

The great majority of stationary internal-combustion engines serve simply as 
prime movers operating single machines or groups of machines by means of main 
shafting, countershafting, and other types of transmission. Direct connection of 
power consumer to prime mover is a special construction of which direct-connected 
electric generators, blowing and pumping engines are well-known examples. 

I. Capital Cost and Cost of Erection 

The question of operating costs touches the designer only indirectly. These costs 
are made up not only of fuel costs, cost of lubrication and attendance, but are strongly 


1 Translator’s Note. This does not apply to the material added describing American machines. 
This was written up simply in the order in which the information was received. 




314 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


influenced also by cost of erection, of maintenance, floor space required, etc. For that 
reason a consideration of the latter factors is therefore not out of place. General 
comparisons of operating costs made up on the usual plan are nearly worthless. 
The items fundamentally concerned in computations of this kind differ so much 
among themselves, that only a close examination of every item in each separate case 
will serve the purpose. It is hoped that the following pages may serve as a reliable 
guide to this end. 

1. Fuel Costs. The statement of fuel consumption should include that used during 
starting up, during stand-by periods, etc. The purchase price of the fuel is understood 
to refer to the actual cost at the place of consumption, that is, freight, loss through 
drying and sifting, and, in the case of gas, the influence of pressure and temperature 
upon meter readings should also be taken into account. Where the load on the plant 
changes considerably, it is usually found that only the total fuel used per operating 
day is known, and not the quantity per H.P. hour. In general, the fuel consumption 
per horse-power changes with the load on the engine, that is, increases or decreases 
with the latter in an inverse ratio. Different engines vary considerably in this respect, 
but the figures in the following table serve as practical average figures for the increase 
of fuel consumption with decrease of load: 


Load approximately. 

Illuminating-gas engine, about. 

Suction gas engine, about. 

Diesel oil engine, about. 


i 

f 

i 

i 


1 of rated load 

10 

20 

35 

60 

90% f greater consumption 

20 

30 

50 

75 

100% { per B.H.P. hour 

10 

20 

30 

55 

80% l than at full load. 


Gas engines in combination with gas producers show a more rapid increase in the 
fuel consumption with decrease in load than either illuminating gas or oil engines. 
For this reason the latter may at certain loads be more efficient than the former when 
at full load the reverse is the case. 

If an engine uses on the average q B.T.U. per B.H.P. hour, and if 1000 B.T.U. 
cost b cents at the place of consumption, the fuel costs per B.H.P. will be 


K b = 


Q 

1000 


b cents. 


( 1 ) 


If the economic efficiency of the installation is 

q ’ 

write 


so that q = 


2545 


we may also 


2545 b 

Tjw 1000 


2.5455 

Tjw 


cents, 


(2) 


Examples. 1 . Suppose that a total efficiency of 18% is guaranteed for a 50-H.P. suction-gas 
installation. Anthracite delivered at the producer house costs $6.80 per ton. The heating value 
of the fuel is 13 500 B.T.U. per pound. We then have 


2545 680X1000 

-= 14 120 B.T.U., b =-= 025 cent 

IS ’ onnAviotnn .uaj ueui, 


2545 ^025 _ 2 ! 545X.025 _ 
.18 1000 .18 


2000X13 500 

.35 cent. 


and 


.18 












STATIONARY ENGINES 


315 


2. Cost of Attendance. The statement often made that a given gas-engine 
installation needs no attendance is of course untrue, for even the smallest size requires 
some attention. It is however certain that gas-engine plants throughout require less 
attendance than steam plants of the same capacity. They require less time, or may be 
served by cheaper labor. The amount of attendance required of course directly depends 
upon the brake horse-power N n of the installation. For the ordinary operating period 
of ten hours per day the hours of labor that must be expended on the installation 
may be approximately expressed by 


W= .25 Wn hours (for illuminating gas and good grade oil engines) . . (3) 

and by W = l.25VNn hours (for anthracite or coke-suction gas plants).(4) 

If the cost of labor, wages of the attendant, is x cents per hour, the cost of labor 
per B.H.P. hour, based on the normal rating of the engine, will then be 


and 


„ .2WNn .025.r ^ ..... . 

K w = —-x = —cents (for illuminating gas and good grade oil engines) . (3a) 

iUiVn V N n 

K w = = cents (for anthracite and coke suction-gas plants) . . . (4a) 

lUiVn V N n 


Assuming an average wage scale of 20 cents per hour, the cost of attendance per 
B.H.P. hour would figure out 


50 

For illuminating-gas engines, K w = - L =. cents 

VN n 


For suction-gas engines, 


iv 2 - 5 

K w = —- cents. . . . 

VNn 


... (5) 

. . . (5a) 


Example. 2. The 50 H.P. plant mentioned in the first example requires a daily attendance 
equal to W = 1.25V 7 50 ~9 hours, which, with the above assumption, makes the cost of labor per 
2.5 

B.H.P. hour equal to K W = V ==-35 cent. 

V 50 


As soon as W exceeds ten or twelve hours, the total time should be distributed 
between two or more men. Taking a 100 H.P. suction-gas plant, for instance, for 
which W = 1.25v / 100 = 12.5 hours, one man could probably do the work ; a 500 H.P. 
plant, on the other hand, should have three men, since W = 1.25V500 = 28. In large 
installations, consisting, say, of n engine units with their producers, the required time 
of attendance may be approximately found from 


or preferably 


W = 1.25VnN n hours 
IF = 1.25n\/nNn hours 


per operating day of the entire installation. 










316 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


If, in cases where W is considerably less than 10-12 hours, the remaining time of 
the attendant cannot be used in other work, his entire wages must of course be 
charged against the plant. 

3. Cost of Lubrication and of Cleaning Material. This item can only be estimated 
on the basis of previous experience. As in every other power plant, the cost of oil 
and of waste is considerably reduced when the oil is recovered and filtered and the 
waste is washed for continued use. In this way the expenditure for these two materials 
may be brought very close to the figures obtaining in steam plants of the same size. 
In medium-sized engines the consumption of lubricating oil is approximately from 
.005 to .008 pint per B.H.P. hour; in favorable cases, however, it may be only .004 
pint and less. 

4. Interest (3 to 4.5%) and Depreciation (7 to 10%), Computed on the First Cost 

(Am) of the Engine together with cost of auxiliaries, and including freight charges, cost of 
foundation and cost of erection, hut excluding the cost of the building. The cost of the 
ground occupied is usually left out of consideration on the assumption that its value 
increases with time. It should, however, be remarked that under certain circumstances 
the cost of acquiring the ground may constitute a large part of the cost of the instal¬ 
lation. Consequently the size of building space required by a given engine installation 
is of some importance as far as operating costs are concerned, although in comparative 
estimates usually only the costs of buildings and foundations appear. The latter of 
course increase with the floor space covered. 1 

5. Interest (3 to 4^%) and Depreciation (2 to 3%), Computed on the First Cost (. A g ) 
of the Building. Where the building or the room is not owned, the rent is substituted for this 
item. If in the same building other machines besides the engine installation are also in 
operation, in computing the operating costs the engine plant is charged up with a 
proportionate part of the sum A g . 

Placing the interest charge at c%, and the depreciation at a%, and assuming that 
the average power developed by the installation is Ne, the amount chargeable against 
the plant per horse-power per year on this basis will be 

K= 100Ne (C + a) dollar? ’.( r> ) 


1 The following figures give a case in point: 

(a) Horizontal suction-gas engine, 100 H.P., with producer, wet and dry purifiers, etc., two rooms 
covering together 1400 sq.ft. 

(b) Vertical suction-gas engine, 100 H.P., with producer, wet and dry purifiers, tar extractor, etc., 
two rooms covering together 970 sq.ft. 

Difference in the floor space is 430 sq.ft, (approx. 30%) which decreased the cost of the space 
approximately $3750. The building for installation (a) costs approximately 1400 X $1.40 = 11960; for 
installation ( b ) approximately 970 X $1.40 = $1360. Difference in the cost of buildings, excluding 

foundation, is $600. The vertical engine installation is therefore in all about $4350 cheaper than the 

horizontal, all other things being the same. 

The cost of all the machinery, erected and ready for operation, is in both cases $8000. Hence, 
according to the above figures, fully one half of this price may be paid out of the saving in cost of 

ground and buildings resulting from the adoption of plan ( b ). The cost of floor space was $8.70 per 

sq.ft., which is not unusually high. In large cities this often reaches $10 and $12 per sq.ft., and 
although commercial plants are not often installed in the select quarters warranting this price, it 
not infrequently happens that in the proposed erection of plants supplying city blocks such prices 
are encountered. 

[Translator’s Note. The above figures apply of course primarily to German conditions.] 





STATIONARY ENGINES 


317 


where A represents first cost, which may be either A m = first cost of engine and auxiliaries, 
as above noted, or A g = first cost of building. 

If Z is the number of hours the plant is in operation per year, then 


-v _ Am ( c+a ) 

m lOONe Z 


dollars per B.H.P. hour 


(7) 


and 


gg ° iU0V ^ Z~ dollars P er B.H.P. hour.(7a) 


Assuming as usual that 


c= 4% and a =8% for A m 


and 


c = 4% and a = 2% for A g 


and taking Z = 3000 hours per annum, eqs. (7) and (7a) reduce to 


K 


m 


250 N e 


cents per B.H.P. hour 


(8) 


and 


K a = - A° - r cents per B.H.P. hour.(8a) 

500A e 


From these equations it appears that the greater the product of N e and Z, the 
lower the cost per H.P. hour, which shows that both high load factors and long 
operating hours are desirable. The same thing applies with even greater force to the 
fuel costs, since these (see under 1) increase much more rapidly with a decrease in N e 
than either K m or K g . As between the individual effects of the factors N e and Z, it 
would seem better to operate a plant for a short time at high load factor than for 
longer hours under a lower load. It is also better to shut down a plant for a daj at a 
time than for an hour now and then, on account of smaller stand-by losses. 

Finally, an inspection of eqs. (7) and (7a) will also show that the bad effect of 
short operating hours upon K m and K g can in a measure be counteracted by small 
outlay in first cost (A m or A g ). In other words, all other conditions being the same 
the less a given plant is used the lower should be its first cost. It is possible of 








318 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


course to reach the same end by proportionately decreasing the charge (a) for depre¬ 
ciation, but that can only be considered when the interruptions in the operation occur 
at regular intervals, when the sum total of the stand-by periods constitutes a large 
part of the year, and when during this time the plant is maintained in its best 
condition. 

6. Maintenance and Repair of Engines and Auxiliaries (2 to 3% of the first 

cost A m ). The quality of the construction and especially the degree of “hardiness” 
against the normal occurrences of ordinary operation play an important part in the 
determination of this item, a fact which should not be lost sight of, especially when 
judging suction-gas plants. In the latter case, poor scrubbing arrangements, causing a 
rapid scoring of the cylinder, bad vaporizer construction, or rapid wear on producer 
lining and grate, may, under certain circumstances, cause a repair bill which exceeds even 
10 or 15% of the first cost. Under normal conditions, however, a charge of 2 or 3% 
is quite sufficient. Extraordinary occurrences, like accidents due to careless attendance, 
etc., are of course not considered in the figure. 

7. Maintenance and Repair on Buildings (1 to l\% of the first cost A g ). 
Strictly speaking, another item that should be taken into account is insurance. But 
the latter charge is very nominal when reduced to the horse-power basis, and is conse¬ 
quently usually neglected. 

In making estimates of operating costs, items 2 to 7 are very often considered as 
merely incidental expenses, making the cost of fuel the only criterion of economy. 
This method, however, applies only in cases where the cost of fuel is very high 
(illuminating gas, and, in part, also crude oil distillates); where the fuel is cheaper 
or is utilized at very high efficiency it will nearly always be found that the “ inci¬ 
dental” expenses constitute the greater part of the total operating costs. Fuel costs 
alone therefore never give a certain indication regarding the economic status of a 
given plant. Plants high in first cost, or those in which the repair and maintenance 
account is high, or those which require constant close attention, occupy much floor 
space or call for expensive foundations, all of these may under certain conditions 
operate with small economy, although the efficiency at which the fuel is burned may 
be of the highest. J 

Tabic 40, and the diagram, Fig. 470, derived from it, gives a convenient survey 
of all of the items concerned in the computation of operating costs. The figures are 
based upon unit horse-power, because it is easier to compare various sizes and types 
of engines better on this basis than if the yearly cost had been made the unit The 
table has been constructed by the use of the equations given above, the assumptions 
or and A a , as well as for fuel costs and incidental expenses, being based upon the 
best obtainable average figures for modern horizontal installations. In order to simplify 
matters the average capacity N e developed by the plant is put equal to the normal 

Si! nS anC the time of °P eration ha s been assumed at 3000 hours per year 
Where these assumptions do not apply, it is easy to use correction factors. 


edition The daU mav not^ f T ^ * rans P osed wit hout much alteration from the German 
true of the first LT s a r ^efore apply xn all cases to American conditions, and this is especially 
an a^erie wai sclT^T ** fcounts. . The latter have been corrected by assuming 

f i . , ^ ‘ ® cents per hour, but the items referring to cost of oil waste costs of 

varfc' aVe n ° tbeenchanged ^ethe entire set of figures is hypothetical and is’subject to large 



FIRST COSTS AND COSTS OF OPERATION, ILLUMINATING-GAS AND SUCTION-GAS INSTALLATIONS 
Based upon 1 B.H.P. or 1 B.H.P.-hour respectively, and assuming 3000 working hours per year 

Figures for Illuminating-gas Engines in Italics. 


% 


STATIONARY ENGINES 


319 


©i 

IqOO 

3.0 

5.0 

57.0 

67.7 


.270 

^ Vh 

05 

17.7 

N. 

T—< 

175 

°.«s 

3.1 

5.1 

58.1 

69.6 

<£> 

.277 



17.7 

7.5 

r—< 

o 

‘o o 

©1 




uO 


• J 

1^ 

CO 

»o 

iO 

rH 

Co o 

*0 r-H 

GO i© 

»o 


gq 

00 

©1 

p- 

1^ 


f-H 

r-H 

1C 

*0 t© 


Ob 



©1 




Ob 

O 

- 

^ o 

t© 

G^ iO 

d 

Co 

CO 

I- 


C5 

©1 




©1 

100 

59.5 

11.2 

3.6 

5.7 

63.1 

76.4 

.250 

305 

vr^. 

©i 

X 

17.7 

GO 

p^ 

©1 


o 

go 


^ P O ©1 

l o 


o 


GO ^S. 


1^ 


lO 


‘O 

Co 

05 

SP 

CO 

.3- 

05 

p- 

1 

0c 

CO 


O 

u 


CO 
C© i“H 


^ r» 
vpo 


Co 

co 


1^ 

r-H 

GO 


*o u- 

Co ©1 
^ CO 


Cb 


vp 

f-H 


GO 




vf- 

Ob 


tO 

-t 


»© 

CO 



*o 1^ 

*0) ©1 

o 



I© 

e.5 

74 

00 

p^ 

05 

cc 


CO 
Co »-h 


Co 

CO 

00 

SP 

CO 

CO 

o 

f*H 

Ob 

uo 

i-H 

ic 

O Id 

Co r " H 

Ob u- 

6'89 

86.2 

•o 

0* 

lO 

-f 

CO 

O *S«J. 

00 

f-H 

rH 

17.7 

10.6 

05 

CD 


66.5 

16.2 

5.2 

8.2 

71.7 

90.9 

»D 

co 

©1 

CO 

CO 

Co ^ 
p^^ 

05 

©1 

rH 

67.5 

18.7 

Vp ID 

*© GO 

72.9 

H 

05 

Ob 

o 

CO 

CO 

^ vp 
00 ^ 

14.0 


p- 


t'- 




00 


Ob 


05 


u 

O 


I 

O 

c/3 

u 

O 

B 


d 
•— 

o 

£ 


Cb to 

Ob u- 
GO H 


O ©1 


o 

co 

Cb 


h 


t© o 

p^ o 
Cb o 


p- to 
00 ©1 





i-H 


u. 

cj 

H 

-*—■ 

• o 

cc 

. • 

b£ 

• ^2 


Co 

>—l 


o 

r- 


0 

r-+ 

• 

t£ 

P 

0 


m 

cz 

co 

P 


co 

o 

CO 

05 

GO 

LO 

*© 

o 


00 

vp 

CD 


©l 


h M S M 
•— o H « 




I 



Cb ©I 

*o 


05 


o 

Co 

•<*p 




CD 

55 

70 

21 

00 

»D 

P^ 

C5 

05 

Cb 

GO 

o 

00 


to 

r-H 



O 

©1 


o: o 

»© ©1 

l O 

©1 

CO 

to 

P^ 

^-■p 

©4 


P^ 

r—* 

©i 

p^ ©i 


GO 

P^ 

CD 

O 

^H 

GO 

©1 

H 

Ob 

>^1 

r-- 

rH 

*“H 

GO 

>**H 

CO 

©1 

i 

l O ©1 

O) ©J 

P^ 

05 

o 

P^ 

Cb 

VP 

o 

1- 


to 

©i 

N *H 
P- CO 

*D GO 

d 

00 

O 

rH 

GO 

co 

CD 

H 

>~7 


©1 

r-H 

Co 

* 27 

iO 

*D ‘D 

O 


to 

o 

P^ 

•o 


H 


>-« 

to 

Co CO 

Co 05 

GO 

co 

05 

©1 

*-H 

Co 

GO 

f-H 

»o 

co 


to 

©1 

rH 

GO 

©1 

co 


*o >-4 

CO 

p>» 

*-H 

20 

05 

*■© 

VP 

o 

I- 

O 

©1 

. vp 



• 


Ob H 

CO 


l © 

CO 

>^4 

-v 

r-H 


LP 







Cl 


6 

p 

co 

<v 

-d 


<D 

d 

*o 

CO 

-♦-» 

to 

O 

O 

bJD 

.5 

%-» 

d 

lH 

<D 

o, 

o 


u2 
3 - 

26? 

-7=0 

<D 

(-i 

a> 

a 


C3 

• ^ 

• O 


r: ■ o 

d o* 

s»/ 

•-* 

CZ3 

0 
• r-H 

’ -H 

• jg 


•d « g 
c u 5 

H 05 

• ! 

• • 0 

t» 

d 

b£ 

i 

* "ri 

* H-3 

• w 

• d 

• H 


o O.CO 
r ° GO 

"S^S 

anly.. 

only. 

engin 


ci 

P 


CO 

O 

0 


w 

c3 

bfi 

i 

P 

O 

a 

d 

co 


>3 

CO 

O 

o 


15 10 13 

H-3 <D H-> 

O O 

■p.S -p 

d ^ d 

O O 


to '*© 
d —^ 
.2^ 


o r*-’ o ^ 

; ‘-r-; £h H d ^ 

cr 2 cr cu d 

co o co o 

“* ’ TS 

<-* 

d 
0 


d 

"O 

r+ 
»—< 

D 

o 


*0 

d 

d 


.0 
/-n P 
O ;d 


d P 
di 

2 72 

.£ c 


C CO 
bJO £ d 

d 2 m 


CO 



©I CO 


H tO 


o 


p- oo 


05 O 


©I 


CO 





























































































320 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


___ j qj 

CD 0) 

I 33 

£ O be 

<u *1 —; 

5 


3 . 

l!; 

g? Sl r 
s «S: 




,‘C 

(h JO 
3 3 
O J 


CX S *> 


J.-S.S 
>> 


g 03 b? CO h 

o -t-s S-h- 
^ «.5 o : 


O 
> 


3oS| 

a CO ^ 
C fl w 

o o 


w ® 

s« 


c e 

8 L* 

*•§ 

Wx 


S>a 


.S '. 510 


S3 


© 


„ Jill 
0^0 © 


o o 




.9 • ©,c 
- :'Eb § 
$ ■ g 

GO . rr Ch 

.S -3.2 

§o :•£ a 

| ; «f 45 
■g B 3 3 

3 « & & 
“ bJrOTJ 
© G G G 
0^33 

0*2*2 
S f^o o 

| Its 18 

l aa 


ex • 

w : 

pq 3 


Sex 

*G i * 
G n 1 : 1 


3 CX 
-c 


HH S 


3 .5 

<D „ 

ft ® c 

M G o 

w ste 

G ej 
© S 


ri 


Is'i 


c °. 2 


J3 g^. 

«s 


exx 
W> . 
« 3 " 

v* 3 ‘ 


JS g&S 
» § 


c3 - 


® ! 


3 


43 

a s -is 

« S,n 

•o Mg 
’o be - ” 

« .S 3 

^ ^ “ 
3 C G 

g g-J 

M go 
• = 3 

£ ~® 


53 2 *■« 

a< o a 9? 

« s 2 
S M:g£g 

ft — fipco 
33 

3 0 0 «? 

2 §glr 

*rS .to 

° a aS 8 

to 3 

5 g g ° ° 
G c-S-S 

HH O O O C 

£H oooo 


.3 O 

O 

G O « 


a • « i 


a ® 

•rr wj 

23 

a> .2 

ol 


^ a 

- .3 O 

§ 9> J £> 


F SO 


Mf § 

i §> 

£i> 

a •- §. 2 ~ 

S -5 _rv3 -s _r 

— O | 

oqO 


mg 3 

c o> 
§> 


i 

Ol 

.0/5 

.022 

.036 

.175 

.005 

.079 

.075 

.095 

■ 4*4 

.641 

14.8 

.83 

1.26 

.28 

1 674 

24.8 

921 

69.6 

IQ 

3- <35 o0 OO 05 in 

«3 (0 00 O *d rHOgOO 00 35 <N 00 <N <N w 3t . 

>~, <M Og 35 <to 00^35 CO .... I. -hO 

<=5 o O HH o 00:0 -4cc 2 ^ ^ ® 


<c n hh uo ^ nco <s cc § ®«s *8 

N Cl " 3 - O OO CC 00 o ^T 35 .... •>• u n 00 o 

CO C Cl O OOHH vt-o 13 hH ^ ^ 


<to rf Hi 13 °0 13 (N SO 00 35 00 C3 M O ^ 5SI 

hh(M M (O »» 0 O ^Cl .... QOwhOo 

CO <C Cl C OQh ^Ih 13 HH ^ ^ 

100 

0 00- Cl — © n-. 

bnfH O O ’h 35 *3 13 95 © Cl 05 CC CO lO .00 . 

h Cl *3 *3 N. 00 ®D O NN .... 00 in ® —, 

CO <C Cl O O O hh v(-tH. O rn ’ Cl ■ U 

80 

d HH C'l >3 _ gn 

Oo 00 <to O ^ CCfHOO *3 0 005hJ«cC C .Ht 

Hh Cl *3 00 C5 eg) 0 05 Cl 3S ^ hh _ 

OO OCI<COC!hh vj-30 0 HH ^H Cl ?- 


d -H d O — O m 

05 05 O O *3 ©S 00 CO O 35 35 Hf CC Hf ®00 . 

■~H Ci C C bn 35 Oo O C O .... 0 

O O <CC0COC>HH *300 Ohh 'd'S 

£ 

C HH *3 *3 N. 05 C O <35 0 CO 0 * CC os ®M 

S) n co ci ?h 05 35 hh h_ q .... os — 31 _ 

9 ^9 5: - 

o 

hhtJ< HH^SQjdCO Nffl tHfflSm 80^*^ 

3| re iHLOCloOCi-H CO-f .... O.ndX 

| 9 9 9 9 - 9 - s: - d ^ s 

40 

^ g g § 5 8^2 §s 9®^9 8 ®3 5 

9 9 9 9 9 hoh — X r-H ^ 

so 

HH-05 j3iO*3»O<30d g|^ rf ®U5M d C) S 10 

9 9 9 r 9 79 rt — 2 - d ^HH ^ 

o 

§ 5 § 2 £ 8 §I 2 §2 -2SS 3 9 

2 0 2 h O -O- ■_; »— „ R 

Ol 1 

fH tH O O ®l 0*3*3 ^ Ch Ih. 2 § CO d ® os ^ 

s s ^ 00H4. ■ - 82 s 

20 

11 s 1 § 2 ia ss -nr s “§ ° 

f-J r-- £1 ^ ^ 

iH j 

03 35 *3dco on—os — 

g‘ | 3 & 1 28 2 

9 0 r 9 r ~9~ v 2 -- ^2 S 

*! 

|| g § g ggg IS 3 2 ^ 0 

lO 

S § 1 ? 2 - «co Ci^gg ? 0 ^ 

!-<•••• ci ci d 1:0 d 00 


G 58 


£ 


dS 


_ - CO £ -H 

I £ g S £ 
9 rt £^2 

liT*: 

^ 3 ^ 53 ° 

^ o 

: O 9 

r s*g 


§"l^§ 

Mo 03 

SfSls 

2 c e3 tc .3 

cS '”' c £ 9 
•S m ® 
gen* ^"5^ 

|w’l|| 

sa!‘ g -9 

«« 3^ 

«° p 30 

.9 ® ► 

ft^OS 
3 o-“ O 

« x’S 

3 d c S ‘S 

•S5 * $ * 

9 m” § 
Q.'O.S G-S 
«_ g.&-2 a 

o § J 
co ^ o c 
C O 

^>#o ®‘C 

3 m 

£ 3 aj O «_ 

X^-9 o ° 

£3.2 « 

a*- 1 G3 ® a 
S-./R 8 >» 

als'Sjj 

^ y ai 

|”siM 

+= tO (C •-* 

g«o® 
„ r ft a.^^-3 

■gS £ « “ 

gg , 8 !e 2 | 

•Jl-r 

O «2 O W H 3 

- g> 

o t'^. 2* .a 

* Is! &1 

ai-so^s 

.. “ G 

N s© 

— £ 3 0 -TV 

» „!2 G 

03 V 




^gai 














































S 10 /S 20 25 30 35 40 4S SO 30 70 30 OO /OO /2S 'SO /7S 


STATIONARY ENGINES 


321 





3 to /5 20 2S JO J5 40 4S SO 60 70 30 90 /OO /2S /SO /7S 200 

a h r 

Fig. 470.—First Costs and Operating Costs of Illuminating and Suction-Gas Installations, (from Table 40). 






























































































































































































































































































































































































































































































































































































322 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


A close study of the diagram, Fig. 470, brings out many things that are not so 
apparent from a survey of the table. Among these may be mentioned the following: 

(a) Rapid increase of both first cost and operating costs with decrease in rated 
horse-power. (Unfavorable to the operation of small plants.) 

(b) The small ratio of incidental expenses (curves h and k) to total operating 
costs (curve m) in the case of the illuminating-gas plants (points to the latter as 
well suited for small capacity and intermittent operation). 

(c) The great preponderance of the incidental expenses (curve 1) over the fuel cost 
(curve g) in the case of the suction-gas plant. 

(d) The rapid increase of the first cost of gas producers (curve b as compared with 
the total first cost of the entire installation (curve e), for capacities less than 30 B.H.P. 

(e) The approach of the curves m and n, showing total operating costs for illuminating- 
gas and suction-gas plants respectively, at the left-hand side of the diagram. 

(/) The considerable proportion that the cost of buildings and foundations bears to 
the total first cost of horizontal machines (distance between curves c and e, or a and d). 

In most comparisons of operating costs, hydraulic power is generally, without 
much consideration, assumed to be the cheapest. As a matter of fact, however, this 
assumption is justified only in cases where this power can be made available with 
simple means and at no great expense, and where the supply of water is uniform 
enough to admit of the elimination of costly reserve machinery. Where these conditions 
do not obtain, it may be found that the first cost of the installation, and the operating 
costs resulting therefrom are considerably higher than the same items for gas or steam 
plants. As an example of this fact, the following figures were taken from a report on 
a 1000-H.P. turbine installation: 1 

Quantity of water available, 141 cu.ft. per sec., under a head of 8.20 ft. 


3 turbines, each 350 H.P. cost. $18750 

3 relay steam engines and boilers, cost. 60000 

Total cost, including generators, buildings, etc. 275000 

First cost per B.H.P., at turbine shafts. 140 00 

First cost per B.H.P., at switchboard. 212 50 

Operating cost per B.H.P. hour, based on general output. 1.35 cents 


The question of the relative superiority of steam and gas as motive power has led 
to a war of words ” between the manufacturers, which has been merely intensified by 
the introduction of the suction-gas plant. With regard to this controversy, the author 
desires to simply note the following: 

In conjunction with good gas producers, and assuming average fuel prices, the gas 

i* S Wlth certaint y able . to compete with steam plants up to 4 or 500 H P 

With cheap gas (less than 70 cents per 1000 cu.ft.) the illuminating-gas engine is 
economically superior to the steam engine up to 50 H.P. Stationary oil engines usually 

en er e e c ^ en illuminating gas cannot be had. An exception to this is the 

lesel engine (and possibly also the Haselwander engine) which, if cheap oils dess than 
0 cents per gallon) are available, remains a serious competitor to steam also for the 
arger sizes and which, where the operating hours are short or where there are frequent 
interruptions, may even be preferable to a suction-gas installation. 

flnnwif 1lera1, S f iS economicall y superior to steam whenever the power demand is 
fluctuating or intermittent and where rapid and. easy erection as well as clean 


1 Suddeutsche Bauzeitung, 1905 , p. 74. 









STATIONARY ENGINES 


323 


operation without molestation of any kind are a consideration. Of course where steam 
power cannot be used on account of government restrictions or other reasons, or where 
on this account its use is rendered difficult, gas power is a very welcome substitute. 

In some of the special types of construction, as direct-connected units for instance, 
the ability of the internal-combustion engine to compete with the steam engine is even 
more pronounced than in the case of simple mill engines, but local conditions have 
a great influence in this respect. The following figures, however, can be obtained with 
certainty, provided the installation is in good shape and the conditions are normal. 

In direct-connected generator units: 

30 to 26.5 cu.ft. of illuminating gas per K.W. hour; 

2.6 to 1.9 lbs. of anthracite per K.W. hour; 

3.3 to 2.6 lbs. of coke per K.W. hour; 

or 1 cu.ft. of illuminating gas will generate from 33.5 to 38.0 watt hours; 

1 lb. of anthracite will generate from 380 to 530 watt hours; 
and 1 lb. of coke will generate from 300 to 390 watt hours. 

In pumping engines: 

1 cu.ft. of illuminating gas will show a duty of from 72-83000 ft.-lbs. ] 

1 lb. of coke will show a duty of from 1000000-1250000 ft.-lbs. [in water lifted. 

1 gal. gasoline will show a duty of from 7775000-8000000 ft.-lbs. J 

These values are obtained when the load factor is high, that is, when the plant 
is operated near normal rating, and refer to standard conditions of fuel. Losses 
through starting, stand-by losses, and belt losses have been taken into account. In 
making guarantees the very strong influence that the load factor has upon plant 
economy must not be left out of consideration. No matter how carefully the plant 
may otherwise be constructed and operated, leaving out of the computation the actual 
average load on the installation may result in complete failure in meeting the guar¬ 
antees made. If therefore the fuel consumption only is the deciding factor in the 
acceptance, the guarantee should be based upon the consumption at or near normal rating. 

In acceptance trials or other tests it frequently happens that differences of 
opinion crop up as to the allowances that should be made for unavoidable fuel 
losses (fuel used until all conditions become constant, fuel not burned at end of 
trial, efficiency of transmission of power consumers, etc.). For that reason it is 
well to have all the conditions thoroughly understood at the time of purchase, or 
at the very latest just before the beginning of the trial. As far as the engine builder 
is concerned, the thing of most importance to him is probably the accurate determi¬ 
nation of the efficiency of the power consumers to be operated (generators, pumps, 
blowing cylinders, etc.), especially since the manufacturers of these machines are very 
prone to give out these efficiencies more or less too high. The power loss in belting 
also is in many acceptance trials a factor of importance. For lack of a‘better way, 
this loss is generally estimated, and it is consequently found that in plants of approxi¬ 
mately equal capacity the allowances vary between 2 and 10% of the engine capacity. 
In reality this loss is probably on the average 4 or 5%; in one definite case, however, 
tests made in the laboratory of a German technical school on a 20 H.P. engine showed 
this loss to be 10% of the total power transmitted. In view of such uncertainties, 
to say nothing of the unavoidable errors of observation, the leeway of 5% which is 
commonly allowed on all guarantee figures is certainly justified. 


324 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 

II. Types of German Engines 

1. Gasmotoren-Fabrik Deutz in Coln-Deutz. (See also Plates VIII to XI.) 


Exhaust 




r - 


Figs. 471-473.—Gas Engine, Model D2 Rated H.P.= 
4 at 200 r.p.m. Net weight including Fly-wheel 
2040 lbs. 


Combined mixing and inlet valve a is automatic, 
exhaust valve b and ignition cock c are operated by 
cam d and rods b' and c' respectively. The latter by 
means of the rod e also actuates the pendulum govern¬ 
or attached to the oscillating slide f. The horizontal 
pick blade of the governor operates the gas valve g. 
A wrought-iron hot tube, heated by the Bunsen 
burner h' is placed in the chimney h. Air is admitted 
to the inlet valve through the inlet muffler and the 
base of the engine. Bottom of cylinder is water- 
cooled. 


Table 41 

DIMENSIONS AND WEIGHTS OF MODEL D 2 ENGINES 


N n in B.H.P. 


R.P.M. 


Length *. ft 

Width f. “ 

Height!. “ 

Net weight.lbs. 

Total weight. “ 


t Along the shaft. 

t To center of shaft. Total height is greater by one-half wheel diameter. 


4 

1 

2 

3 

4 

240 

230 

210 

210 

200 

2.44 

3.20 

3.50 

3.80 

4.27 

2.01 

2.38 

2.53 

2.84 

3.17 

3.20 

3.50 

3.60 

3.66 

3.66 

750 

1080 

1500 

1930 

2640 

965 

1430 

1830 

2330 

3120 








































































































































































































































0 

+3 

g 

u 

G 

Cu 

0 

co 

>> 

J3 

T3 

i 

43 

G 

- 

0 

Gn 

o 

0) 

t- 

G 


0 

j> 

*g 

> 

co 

G 

fciD 

T3 

G 

G 


0 

> 

+3 

CO 

G 

G 

"S 

0 


o 


a? 

r-« 

H 


”0 

.> 



0 


CO 

a> 

tn 


"G 

c 

G 


rO 


V 

e 


CO 

O 

s-» 

u. 

O 


CO 


a? 




0 


CO 

£ 

G 

0 


CD 

>> 

0 

G 


-f3 


o 

u 

O 

G 

- 

0 

•*-4 

3 

G 

cr 

G 

CO 

0 

.a 

0 

-43 

G 

t-4 

> 

<d 

A 

-4-3 

CO 

0 

o 

bo 

0 

G« 

O 



A 


G 

b C 

-43 

O 

bJO 

G 

(4-4 

.3 

*4-< 

o 

■43 

0 

*G 

"g 


'G 

-43 


— 

G 

G 

<—i 

<1 

3 

0 

o 

o 

o 


0 

0 

>> 

0 

-G 

^3 

.G 

-g 

p>> 

■4-5 


G 

G 

s 

<—< 

"G 

-4-3 

o 

- 

+3 

*43 


, 

o 

o 

^G 

pH 

0) 



O 

• <-4 

fcG 


G 

■4-3 

0 

P 

0 

G 

0 

■4-3 

GQ 

G 

r4 

-43 

*■4 

<D 

G- 

O 

(1) 

c/T 

3 

0 

p^ 

* 


CO 

p * 

<■4 

• r4 


i 

^0 

^4 

£T 

K*> 

T3 

<— 

Q 

G 

• pH 

qi! 

G 

• f—i 

CO 


-4 

p3 

-43 

^G 

>> 

bJO 

G 

'"G 

G 

.G 

*bi) 

G 

t£ 

CO 

H 

PH 

G 

a; 

Ph 

• I—1 

pH 

G 

r4 

'S 

CD 


.2 

0 

- 

G 


4-3 

> 

o 

co 

-43 

(4-H 

M 

bX) 

o 

i 

G 

hH 


a? 

r-* 

*—< 

pH 

a 


> 

-43 

o 

o 

bJO 

0 

s-, 

r* 

CO 

c 

G 

43 
• -4 

V. 

o 

o 

G 

-43 

.X 


G 

0) 

H 

44 

0 

— 

G 

O 

3 

G 

K. 

0 

P 

H 

.CO 

CO 

p 

o 

+3 

G 

G 

p 

«4-4 

G 

bX) 

G 

O 

*4-4 


o 

s- 

o 



fc4 

a? 

,G 

-*3 

• *"H 
0 







P 


5 



sat 

IJ 

K 



Al 

cz_=r 


in—T 

TTs 

aidm- 

(p: _ -'L 1 T 

l" 1 

_ W _ 

— 

































































































































































































































































































































































































































































































































































































































326 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


Table 42 

DIMENSIONS AND WEIGHTS OF MODEL H 3 ENGINES 


N n in B.H.P. 

6 

8 

10 

12 

16 

R.P.M. 

300 

300 

300 

300 

300 

Length. 

. ft. 

4.27 

4.35 

4.76 

4.76 

5.09 

Width. 

. “ 

4.30 

4.87 

5.42 

5.42 

5.88 

Height. 

. “ 

5.75 

6.20 

6.24 

6.24 

6.86 

Net weight*. 


2530 

3400 

4050 

4290 

4840 

Shipping weight. 


3080 

4380 

4840 

5170 

5720 


* These and all the weights given in what follows refer to mill engines. 

For electric-lighting service (d=yb) the engines are heavier by from 15-20%. 



Figs. 480 and 481.—Gas Engine (Old Type G 6 III) for Capacities from 35-150 B.H.P. 

v ?°V n l et Valve a and exhaust valve & are placed on the under side of the 
c\ inc er lea , c is the gas valve, d the starting valve. All valves are operated by 
cams, the valve levers are marked a', b', and c' respectively. The make-and-break 
ignition is placed in the axis of the cylinder. Speed is regulated by hit-and-miss or 
quality governing For capacities from 50 to 150 H.P., the firm has lately introduced 
a oew mo e 1 ^ w ich, as far as frame and cylinder are concerned, is the same as 
Ob ill, but has the arrangement shown in Plates IX and X regarding valves, valve 
gear and governing. The dimensions given in the following table will serve approxi¬ 
mately for both types: 

























































































































PLATE VIII 





Legend: 

a = Gas pipe for ignition flame: 

6 = Cooling water supply pipe. 
c = Cooling water discharge pipe; 
d = Cylinder drain cock; 
d' = Drain cock for muffler. 


Rated B.H.P. = 12 at 200 r.p.m 


Erecting Plan for an Illuminating Gas Engine (Mod. E 3). Gasmotoren=Fabrik Deutz in Coln=Deutz 
































































































































































































































































































































































































PLATE IX 


Inlet valve lever a, through the link 
6, operates the lever c of the gas and 
air valve located in the inlet valve hous¬ 
ing Between lever c and bridge c' there 
is interposed a roller whose position is 
controlled by the governor through d. 
The position of d determines the arm 
ratio of lever c, and thus changes the lift 
of the gas and air valve. 



Part III shows following ad¬ 
ditional details: 

Figs. 167 and 168: Piston and 
Rod 

Figs. 193 and 194: Crank¬ 
shaft 

Figs. 217 and 218: Connec¬ 
ting-rod. 

For latest method of Govern¬ 
ing see Fig. 329. 


Assembly Drawing, 600 H.P. Double=acting 4=cycle Twin Engine, Gasmotoren=Fabrik Deutz, Coln-Deutz. 



















































































































































































































































































































































































































































































































































































































































































































































































































































, 























































































































































PLATE X 



Assembly Drawing, 600 H.P. Double=acting 4=cycle Twin Engine. Gasmotoren=Fabrik Deutz, Coln=Deutz. 

For Vertical Cross-section see Plate IX. 









































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































PLATE XI 






- -~-=o 

•>s ; 

a j. i 

fv 5 i 

^ 5 <3 

1 

rfi hi 

_X w- P - 

p=--=^—~3 




/<et/ to Colors. 


| Oa-s PLpe 


Air Pipe 


x/icic/st Pipe 


Cooling lA/ater 
Supp >2 if Pipe 


^Cooling UVader 
wESADI’S charge Pipe 



Ventilating Pipe 


\CompressecLAtr 
I Supplif Pipe. 


Cl Co rnp rested Air 
2 g ,Pischa rye Pipe. 


Erecting and Piping Plan 
for a 

Double-acting Twin 4-cycle Suction-Gas Engine. 
Gasmotorenfabrik Deutz A.-G., Koln-Deutz. 

Rated Brake Horse-power 600 at 150 r.p.m. per min. 

(Cylinder Diameter 25.2 Inches. Stroke 30.7 Inches.) 


Assembly Drawings of Engine on Elates IX and X 



































































































































































































































































































































































































































STATIONARY ENGINES 


327 


Table 43 

DIMENSIONS AND WEIGHTS OF MODEL G 6 III ENGINES 


Nn in B.H.P. 

35 

40 

50 

60 

70 

80 

100 

125 

150 

R.P.M. 

190 

190 

190 

190 

180 

180 

170 

160 

140 

(A 

3.70 

3.70 

4.14 

4.19 

4.72 

4.72 

5.43 

5.96 

7.17 


B * 

.79 

.82 

.95 

1.03 

1.15 

1.21 

1.31 

1.47 

1.80 

Dimensions in feet.■ 

C 

1.39 

1.39 

1.56 

1.56 

1.64 

1.64 

1.80 

1.97 

2.38 


c 1 

2.62 

2.62 

2.79 

2.79 

2.95 

2.95 

3.12 

3.28 

3.61 


At 

7.22 

7.55 

7.88 

8.20 

9.18 

9.18 

9.52 

10.30 

11.50 



10.35 

10 35 

11.30 

11.30 

11.85 

11.85 

14.00 

14.95 

16.15 

Net weight. 

.lbs. 

15400 

19800 

21200 

24200 

28 600 

30800 

35200 




Up to 80 B.H.P. inclusive, with or without outboard bearing; in the latter case there are two fly-wheels. 
♦For mill and electric lighting purposes, 
f For mill purposes. 


Since the beginning of 1904 the Deutz Co. use the double-acting 4-cycle engine 
for all powers above 150 B.H.P. The main features of the design of this machine 
are shown in Plates IX and X. Greater capacities are obtained by the combination 
of cylinders side by side or tandem, or both, resulting in twin or tandem, or twin- 
tandem engines. In this way capacities up to 6000 B.H.P. may be reached. 



Figs. 482 and 483.—Erecting Plan for a Deutz Gasoline Engine, Model E. 

The room for gasoline storage is separated from the engine room by a solid wall (following the insurance 
regulations required by the Union of German Fire Insurance Societies). The gasoline tank shou d never je 
located higher than the carbureter to prevent syphoning. 

Operating Results, (a) Illuminating Gas Engine, horizontal cylinder, 50 B.H.P., 
D = 15", £ = 22.85"; ratio of compression £ = 5.92; ignition by hot tube; speed regu- 



























































































































328 CONSTRUCTION, ERECTION, TESTS OF INTERNALrCOMBUSTION ENGINES 


lation by changing quality of mixture. Tests made by Prof. E. Meyer, 1 August, 
1908, showed the following results: 


Table 44 


Test 

No. 

R.P.M. 

n 


I.H.P. 

Mean 
Effective 
Pressure, 
fromiy i 

Vi 

lbs. per sq.in. 

Nr — 
Ni- 

N e 

H.P. 

Mean 

Friction 

T}m = 

N e + 
Ni 

% 

Gas Consumption 
per 

Heat 
Lost in 
Cooling 
Water 
per Cu.ft. 
of Gas. 

B.T.U. 

Ex¬ 

B.H.P. 

Ne 

Total, 

Ni+ 

Net* 

Ni 

Resistance, 

TV 

lbs. per sq.in. 

B.H.P.- 
hour. 

cu.ft. 

I.H.P.- 

hour. 

cu.ft. 

haust 

Temp. 

Deg. F. 

1 

198.8 

64.2 

77.0 

72 A 

70.5 

8.2 

8.10 

89 

15.65 

13.90 

210.15 

929 

2 

204.6 

53.7 

67.2 

62.5 

59.2 

8.8 

7.95 

86 

17.75 

15.40 

199.00 

850 

3 

206.3 

42.0 

56.5 

51.8 

48.7 

9.8 

9.25 

81 

20.45 

16.60 

194.50 

738 

4 

207.9 

30.3 

44.5 

39.5 

36.8 

9.2 

8.67 

76 

28.90 

21.10 

186.50 

616 

5 

206.0 

17 *7 

33.8 

29.0 

27.3 

11.3 

10.70 

61 

44.10 

26.90 

170.00 

435 

6 

200.6 

0 

17.0 

12.3 

11.9 

12.3 

12.00 

0 

417.0 per 
hour 

33.90 


287 


* See footnote p. 342. 


The gas consumption is referred to 32° F., 760 
mm. barometer, and a heating value of 560 B.T.U. 
per cu.ft. 

The compression pressure was 131 lbs. and the 
explosion pressure at maximum load from 312 to 
355 lbs. per sq.in. Diagrams obtained at maximum 
load are shown in Fig. 484. The rounded tip of 
the combustion line indicated strong after-burning up 
to 1/15 of the expansion stroke. 

(6) In 1895, 0. Korner obtained the following 
values with two of the smaller horizontal machines: 



Normal Rating, B.H.P. 

R.P.M. 


8 

222 

25 

201.2 

B.H.P. 


9.86 

9Q 7ft 

Heating Value of Gas, B.T.U. per cu.-ft_ 


566 

560 

Gas Consumption, cu.ft. per B.H.P.- hour. . . ] 

i at 32° and 760 mm. j 

18.50 

17.20 

Gas Consumption, at no load, cu.-ft. per hour. J 

’ including ignit. flame \ 

36.50 

91.50 


(c) Pressure Producer Gas Engine , 160 B.H.P., 2 cylinder opposed, each of 
dia. 20.5", stroke 30", ratio of compression e=4.85. Electric ignition, hit-and-miss 
regulation. Each cylinder has two exhaust valves, the discs of which are cooled 
by spraying from below with water. Tests were conducted by E. Meyer in April, 
1896, on an engine in a water works. 2 

The compression pressure was 115.5 lbs., the explosion pressure, on account of 
slow combustion, was but little higher than this. Heating value of the coke 12,995 
B.T.U. per lb., of the producer gas was on the average 135 B.T.U. per cu.ft., referred 
to 32° F. and 760 mm. barometer. 































STATIONARY ENGINES 


329 


Table 45 


R.p.m. 

n 

Mean 

Effective 

Pressure, 

Pi 

lbs. per sq.in. 

I.H.P. 

H.P. in 
Water 
Lifted, 

N w 

B.H.P. 

Com¬ 

puted, 

N e 

Coke Consumption on Basis of 

Cooling 

Water 

per 

I.H.P.- 

hour. 

cu.ft. 

Water 

Vapor 

per 

Pound 

Pro¬ 

ducer 

Coke. 

lbs. 

Ni 

N w 

Ne 

Pro¬ 

ducer. 

lbs. 

Boiler. 

lbs. 

Pro¬ 

ducer, 

lbs. 

Boiler. 

lbs. 

Pro¬ 

ducer. 

lbs. 

Boiler. 

lbs. 

141.4 

48.00 

169.0 

121.5 

142.0 

1.27 

.132 

1.785 

.193 

1.51 

.152 

.61 

.614 






1.402 

1.978 

1.662 



137.75 

51.20 

176.0 

130.0 

149.6 

1.28 | 

1.42 

1.735 | 

.193 

1.49 | 

.152 

.81 

.642 






1.422 

1.928 

1.642 




One of the older types of pressure producer gas engines of 100 H.P., tested in 
1897 by O. Kohler, gave the following figures: 

Fuel used in the producer, German anthracite, heating value 14400 B.T.U. per lb. 
Fuel used in the boiler, gas coke. 

Mean heating value of gas at 32° F. and 760 mm. bar, 146 B.T.U. per cu.ft. 


Mean R.P.M., 160.93, B.H.P. 114.1. 

Fuel consumption per B.H.P., per hour in producer.966 lbs. 

“ “ “ “ “ in boiler.126 “ 

“ “ “ “ “ total. 1.092 “ 


(d) Results of Tests on several Deutz Suction Gas Installations. Test by F. 
Kapeller, Director of the Industrial School, Niirnberg, 1902. 

Engine rating, 20 B.H.P. 

B.H.P. on test, 23.10. 

Consumption of anthracite per B.H.P.-hour, .995 lbs. 

Coal, from Langenbrahm mine, had a heating value of 14940 B.T.U. per 
lb., and contained 2.71% of ash. 


At about the same time tests made by A. Stauss in the mechanical laboratory 
of the technical school at Stuttgart gave the following results: 1 


Normal rating, 70 K.P. 

Cylinder diameter, 16.10" 

Stroke, 23.6". 

R.p.m., approx., 182. 

Mean piston speed, 11.95 ft. per sec. 
Compression press., 9 at. = 128 lbs. per sq.in. 


Langenbrahm anthracite, size $—$"• 

Heating value of anthracite, 13 925 B.T.U. per lb. 
Composition of coal: 

C = 84.60%, H = 3.49% 

O + N = 3.49%, S= .92% 

Ash= 5.82%, Water=1.20% 


1 See Journal fur Gasbeleuchtung, 1902, Heft 29, and following. 


































330 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


Time of reading. 11:30-2:44 

Mean effective pressure. 68 0 

I-H.P. 72.25 

B-H.P. 64.10 

Mechanical efficiency, %. 88.65 

Consumption of anthracite per I.H.P. hour, lbs. net. .897 

Consumption of anthracite per B.H.P. hour, lbs. net. 1.010 

Consumption of water (steam) per I.H.P. hour, lbs. 0.492 

Consumption of water (steam) per B.H.P. hour, lbs. .556 

Water per lb. of anthracite, lbs. 1 205 

Water per I.H.P. hour, lbs. 16.25 

Water per B.H.P. hour, lbs. 18.35 

Cooling water per I.H.P. hour, lbs. 43.10 

Cooling water per B.H.P. hour, lbs. 48.70 

Total water consumption per I.H.P. hour, lbs. 59.35 

Total water consumption per B.H.P. hour, lbs. 67.05 

Outlet temperature of scrubber water, ° F. 117.8 

Outlet temperature of cooling water, ° F. 136.4 

Efficiency of producer, %. 86.36 

Economic efficiency of entire plant, %. 17.99 

Heat lost in scrubber, % . 8.84 

Heat lost in cooling water, %. 29.79 

Heat lost in exhaust , % . 36.28 

Composition of producer gas, volume %: 

CO. 22 0 

. 17.0 

CH 4 . 1.8 

Cnlln. q 

C °2. 5.8 

. * .7 

*. 52.7 

Composition of the exhaust gases: 

C0 2 = 5.6%, 0 2 = 13.6%, N = 80.5%, combustible (taken as CO) = 
mains: 3-5" water below atmosphere. 


2:44-6:48 

66.6 

71.60 

63.90 
89.35 

.787 
.892 
.494 
.550 
1.372 
16.48 
18.55 

32.90 
36.70 
49.38 
55.25 

109.2 

154.4 

86.82 

20.61 
8.93 

31.14 

32.61 

23.3 

17.4 
2.0 
0 

5.5 
.5 
51.3 


.3%. Pressure in gas 



-20 at 


-10 


The explosion pressures varied between 200 
and 230 lbs per sq.in. In all of the tests the 
combustion was so slow that the diagrams, see 
Fig. 485, closely resembled those of a constant 
pressure engine (Diesel for example). 

Tests made by R. Mathot, March, 1904, 
on a 60 H.P. engine having 16.5" cylinder 
diameter and IS. 9" stroke, gave the following 
main results: 


Load. 


I.H.P. 

B.H.P. 


Maximum suction pressure, pounds per square inch .... 

Relative increase of speed (full load = l .0). 

Consumption of anthracite per I.H.P. hour, lbs. net .... 
Consumption of anthracite per B.H.P. hour, lbs. net. 


Full. 

Half. 

No Load. 

4.5 

4.5 

_ 

188.66 

195.5 

199.0 

73.0 

44.3 

11.8 

75.9 

44.35 

10.83 

64.20 

33.35 

_ 

84.6 

75.0 

_ 

170.0 

121.0 

92.3 

383.0 

249.0 

213.0 

24.1 

19.8 

0 

-4.26 

-6.52 

-8.52 

1.0 

1.035 

1.052 

.68 

.876 

_ 

.798 

1.170 

_ 

24.3 

16.0 

_ 


ash 


The coal used 
7.33%, water 


came from 
2.69%, 


the 


heating 


Mossbach mine near 
value 13638 B.T.U. 


Aix-la-Chapelle, size 
per lb. The gas 


-H 

had 















































STATIONARY ENGINES 


331 


heating value of 140 B.T.U. per cu.ft. At full load, the outlet temperature of the 
jacket water from the cylinder head was 109.4°, from the cylinder barrel was 127.4°, 
while the temperature of the vaporizer was 183.2°. 

The following figures are taken from the results of a two-dav test made by 
Prof. Witz and Mr. Mathot in 1894 on a 200 H.P. double acting 4-cycle engine 
(D = 21.2", 3 = 27.5", piston rods 4.72 and 4.33"). 

First Day. Second Day. 


Duration of test, hours. 3 10 

Average revolutions per minute. 151.29 150.20 

B.H.P. 211.8 220.00 

Consumption of anthracite per B.H.P. hour, lbs. .937 .726 

Economic efficiency, entire plant, %. 19.0 24.4 

Total water consumption per B.H.P. hour, lbs. — 77.8 

Of this amount there were used in cylinder jackets. — 45.8 

in piston and rod. — 17.25 

in producer vaporizer. — .65 

in scrubber. — 14.10 

Water vapor per pound of anthracite, lbs. — 1.92 


The temperature of the cooling water was 117.8° leaving the piston and 134.9° 
leaving the cylinder jackets. The gas had a temperature of 557.6° leaving the pro¬ 
ducer and 62.6° after passing the scrubber. The kind of anthracite used (Bonne 
Esperance et Batterie, Herstal) had a heating value of 14580 B.T.U. per lb. When 
banked over night the producer burned 47 lbs. of coal, which is approximately 10% 
of the daily consumption. 

(e) The following table shows average results obtained on. capacity tests made 
by the firm on their new brown-coal installation. The plant was one of the first 
ones constructed and developed 60 H.P. 

Table 46 



Rhenish Brown Coal. 

Styrian 
Brown Coal. 

Test 

I 

Test 

II 

Te3t 

III 

r C. 

23.86% 

1.82 

9.21 

.29 

4.64 

60.18 

30.99% 

2.10 

11.44 

.25 

2.67 

52.55 

29.78% 

2.03 

11.975 

.085 

2.18 

53.95 

44.22% 

3.58 

17.74-!- .52 
.29 
5.86 
27.79 

H . 

Composition of coal as mined jo* 

Ash . 

. Water. 

in pnnl . 

103.00 

35.18 

3283 

. 100.00 

44.78 

4320 

103.03 

43.87 

4257 

100.00 

66.35 

6840 

TTootinfr xtqIiip nf pnfll R T TT DPI' DOllHcl . .. 

ilCatlll^ ValUC LUalj . A . vv . poi . 

8-9% 

23-25 

12-13 

35-38 

123 

18.1 

93-96.5 

5.18 

8% 

25 

15 

40 

140 

25 

86.2 

3.35 

7-8% 

27 

15 

40 

140 

24.5 

82.1 

3.35 

3.85% 

28.70 

13.65 

42.35 

146 

35.3 

about 78.5 
“ 2.23 

A trprnorp PomDfV?ltinTl OT P‘3S \ ( ( ) . . . 

AVcIatc l/Ulllj/Uoitiuu yJi g,ac» i .. 

[ H . 

PomKiKf ihlp in (T £\8 W. . 

Average heating value of gas, B.T.U. per cubic foot . . . 

Vjab ODltlllieu uci puuiiu ui wtuj .. 

\_ias consumeu pei xj.aa.x . uuui , . . 

LOal consumea pci IJ.II-I • nuui, . . 


* Exclusive of sulphur in ash. 




















































332 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


It is stated that these brown-coal producers are capable of gasifying very inferior 
coals containing as high as 60% of water, and that they may be banked for periods 
even exceeding 14 days. The starting up after such periods, it is claimed, does 
not take over 15 minutes. 

Tests were also made with air-dried peat (16.57% water); the peat consumption 
per B.H.P. hour is given as 2.84 lbs. 

One of the new double-generator plants for brown coal (see Fig. 432, p. 283> 
was tested by Prof. E. Meyer in June, 1903, with the following results: 


Test No. 

I 

II 

III 

Duration of test, hours. 

91 

12 

10 

Revolutions per minute. 

181.8 

181.6 

182.7 

Brake H.P. =N e . 

69.50 

69.45 

no load 

Positive ind. mean pressure, p t + lbs. per sq.in. 

67.1 

67.5 

27.80 

Negative ind. mean pressure, pi lbs. per sq.in. 

5.13 

5.03 

6.74 

Net ind. mean pressure, Pi = Pi+~Pi- lbs. per sq.in.. 

61.97 

62.47 

21.06 

I.H.P. (from Pi) =Ni . 

79.50 

80.00 

27.00 

Friction horse-power, N{—N e . 

10.00 

10.55 


Mechanical efficiency, %. 

87.5 

86.9 

_ 

Gas consumption per B.H.P. hour, cu.ft. 

78.4 

78.6 

_ 

Gas consumption per I.H.P. hour, cu.ft. . .. 

68.6 

68.3 

107.5 

Heat value of gas (32° and 760 mm.) B.T.U. per cu.ft 

112.0 

113.0 

101.0 

Heat used in gas, per B.H.P. hour, B.T.U. 

8 770 

8 850 


Heat used in gas, per I.H.P. hour, B.T.U. 

7 650 

7 680 

10850 

Consumption of brown coal per B.H.P. hour, lbs. . . 

1.865 

1.61 


Consumption of brown coal per I.H.P. hour, lbs . .. 

1.63 

1.40 


Heat used in coal per B.H.P. hour, B.T.U. 

14 250 

12 320 


Heat used in coal per I.H.P. hour, B.T.U. 

12 570 

10 680 


Volume of gas per lb. of coal, cu.ft. .. . 

42.1 

48.5 

43.2 

Efficiency of producer, %. 

61.3 

71.6 

56.6 

Total economic efficiency of installation, % .... 

17.75 

20.5 



The brown coal used, Agnesschacht mine, had an average heating value of 7670 
B.T.U./lb. The composition of the gas made is shown in the following table: 


Test No. 

C0 2 , % by vol. 

O, “ . 

I. 

II. 

14.71 

.41 

III. 

14.31 

1.12 

H 2 " . 


17 ^7 

ch 4 


±4.04 

14.08 
1.71 

n 90 

CO “ . 


1 . OO 

12 14 

N 


^4. 

y .oo 

PH A C 



Ot . OO 

oU .45 

Heating value by calculation, B.T.U. per cu.ft. 

100.00 
... 112.0 

100.00 

113.0 

100.00 

101.1 


The recent catalogues of the Deutz Co. contain the following additional tests on 
brown coal suction-gas plants: 


Rated B.H.P. 

Revolutions per minute. 

B.H.P.. 

Coal consumption per B.H.P. hour, lbs. 

Heating value of brown coal briquettes, B.T.U. per lb. 


60 

144.1 180.67 

65.4 98.5 

1.218 1.115 

9200 


160 


167.7 101.1 

1.277 1.338 

9180 


(/) Petroleum and Gasoline Engine, 6 B.H.P., D=6.7", 5 = 10.2", horizontal type. 
Compression pressure appr. 57 lbs. Hit-and-miss speed regulation, and electric igni- 
tion. Tests made by Prof. E. Meyer, 1899, alternately with gasoline and kerosene, 
showed results as per table following: 1 


1 Z. d. V. D. I., 1900, p. 330. 
































STATIONARY ENGINES 


333 


Table 47 


Fuel. 

Gasoline. 

Kerosene. 

Load. 

Full. 

Half. 

No Load. 

Full. 

Half. 

No Load. 

Revolutions per minute. 

242.0 

244.9 

247.3 

240.2 

244.2 

246.7 

Ignitions per minute. 

110.2 

56.6 

19.5 

109.5 

58.4 

12.4 

B.H.P. 

6.93 

3.02 

0 

6.87 

3.02 

0 

Fuel consumption per B H.P. hr., lbs.. . 

.681 

.950 

1.53 per hr. 

.923 

1.271 

1.48 per hr. 


(g) Alcohol Engine (Locomobile), 14 B.H.P., Z> = 8.28", S = 11.80", ratio of com¬ 
pression £==5.91. Speed regulation by changing quantity of charge (taper cam for 
inlet-valve lever), electric ignition. Internal vaporization, that is without special 
vaporizer, starting with gasoline. 

Various tests made by Prof. E. Meyer, in March and May, 1901, showed the 
results given in the following tables. For the first the fuel was 87.2 weight per cent 
alcohol, having a lower heating value of 10440 B.T.U. per lb. 


Table 48 


Test No. 

Load. 

R.p.m. 

n 

Com¬ 

pression, 

Lbs. 

B.H.P. 

Alcohol per 
B.H.P. Hour. 
Lbs. 

Economic 

Efficiency. 

% 

Outlet Temperature of Cooling Water. 

Cylinder 

Head. 

Cylinder 

Jacket. 

Average 
of Both. 

1 

max. 

275.2 

107 

15.68 

.960 

25.3 

208.4 

172.0 

201 .2 

2 

normal 

280.3 

100 

13.91 

1 .00 

24.3 

208.4 

179.5 

208.4 

3 

l 

279.1 

— 

10.42 

1.060 

23.1 

208.4 

170.5 

203.0 

4 

h 

282.7 

53.2 

7.05 

1.218 

20.0 

208.4 

176.0 

208.4 

5 

I 

291.4 

— 

3.64 

1.885 

13.0 

208.4 

172.0 

208.4 

6 

no load 

293.5 

21-15 

— 

5.85 per hr. 

’ 

208.4 

201.2 

208.4 


The consumption of cooling water varied from 21.6 to 19.5 lbs. per B.H.P. hour, 
being small because of the high outlet temperature. 

Tests with alcohol-benzol mixtures made by the same experimenter on the same 
engine gave the following results: 

Table 49 


Kind of Fuel. 

Commercial Alcohol 
86.4 Wt. % Alcohol 
and 13.6 Wt. % 
Water. 

H u = 10 350 B.T.U. 

Mixture: 

9.12 Wt. % Benzol 
and 90.88 Wt. % 
Alcohol. 

H u = 10 980 B.T.U. 

Mixture: 

14.3 Wt. % Benzol 
and 85.7 Wt. % 
Alcohol. 

H u = 11 340 B.T.U. 

Mixture: 20.95 Wt. % Benzol anti 
Wt. % Alcohol. 

H u =11 808 B.T.U. 

79.05 

Test No. 

2 

1 

3 

7 

8 

9 

4 

5 

6 

10 

11 

12 

13 

14 

15 

Load. 

Maxi¬ 

mum. 

Nor¬ 

mal. 

i 

Maxi¬ 

mum. 

Nor¬ 

mal. 

i 

Maxi¬ 

mum. 

Nor¬ 

mal. 

i 

Maxi¬ 

mum. 

Nor¬ 

mal. 

1 

i 

i 

No 

Load. 

R.p.m. 

280 

273.3 

286 

277.5 

278 

286.7 

275.8 

278.3 

287.2 

279.1 

280.2 

286 

287.8 

291.7 

297.3 

B.H.P. 

15.92 

13.60 

7.20 

15.78 

13.87 

7.21 

15.68 

13.90 

7.24 

15.85 

14.00 

10.72 

7.24 

3.72 

— 

Fuel per 















— 

B.H.P. 

.993 

1.033 

1.175 

.942 

.968 

1.16 

.865 

.918 

1.12 

.837 

.859 

.887 

1.17 

1.47 

3.82 

hr., lbs. J 















per hr. 

Econ. eff.,% . 

24.7 

23.8 

20.8 

24.5 

23.8 

19.8 

25.85 

24.3 

20.0 

25.7 

25.0 

24.3 

20.1 

14.7 



(h) In the main trials of alcohol locomobiles, made under the auspices of the 
Deutsche Landwirtschaftliche Gesellschaft in 1902, a Deutz engine gave the results in 































































334 CONSTRUCTION, ERECTION, TESTS OF INTERNALrCOMBUSTION ENGINES 


the following table, compiled from the report of Prof. Eugen Meyer. The machine 
was of the horizontal type, D=8.30", &=11.8", ratio of compression £=8.90, piston 
speed 9.2 ft. per sec. The engine had no special vaporizer, and was regulated by 
throttling the charge. 


Table 50 


Load. 

R.p.m. 

B.H.P. 

Consumption per B.H.P. 
Hour. 

Econo¬ 
mic Effi¬ 
ciency, 
% 

Volume 
of Gharge 
in % of 
Cylinder 
Volume. 

Max¬ 
imum 
Suction 
Pressure, 
lbs. per 
sq.in. 

Com¬ 
pression 
Pressure, 
lbs. per 
sq.in. 

Explo¬ 

sion 

Pressure, 
lbs. per 
sq.in. 

Exhaust 
Temper¬ 
ature, 
Deg. F. 

86.1 Wt. 

% 

Alcohol. 

lbs. 

Cylinder 

Oil. 

lbs. 

Jacket 
W ater, 

lbs. 

Maximum . . 

276.9 

16.55 

.813 

.0057 

1.69 

31.6 

93 

- 2.40 

193.0 

468.0 

>900 

Normal .... 

284.1 

11.93 

.867 

— 

2.10 

29.6 

71 

- 5.23 

136.0 

270.0 

892 

Half. 

292.5 

'6.18 

1.115 

— 

2.60 

22.7 

42 

- 9.08 

76.5 

178.0 

892-897 

No load .... 

298.2 

— 

4.63 

.0136 

— 

— 

19 

-11.50 

36.8 

113.5 

585-625 




per hr. 










The alcohol used was 86.1 w r eight-per-cent, for which Meyer computed a heating 
value of (.861X11 664)-(.139X1080) = 10 043-150 =9893 B.T.U. per lb. 

2. Gebr. Korting A-G in Hannover (See also Plates XII to XV). 

This firm makes the single-acting ^-cycle machine shown in Fig. 486, in sizes 
from 2 to 175 H.P.; from 175 to 350 H.P. the same design is used as a twin engine. 
The double-acting 2-cycle engine shown in Fig. 493, p. 338, is made in sizes from 
350 H.P. up. 

The most important details of the 4-cycle design are set forth in Figs. 487 to 492 
following. With reference to Figs. 490-492, a is the inlet valve operated by cams and 
levers a' and b' respectively. Mixing valve c is automatic, while the starting valve 
d is operated by hand by the lever d'. Electric ignition e is double. Combustion 
space and cylinder are connected only by the comparatively narrow passage /, formed 
by the vertical projection g, which is effectively water-cooled. The idea of this 
construction is to increase the cooling surface for the purpose of raising the com¬ 
pression. At the lowest point of the cylinder there is provided a blowing-off and 
drain cock h. The speed of these engines, as in all machines made by this firm, is 
controlled by throttling the charge—in this case by a butterfly throttle-valve placed 
in the inlet passage between the mixing valve and the inlet valve. 

In the later construction, Figs. 487-489, the vertical projection g has been replaced 
by a horizontal water-cooled partition, again for the purpose of increasing the cooling 
surface and decreasing the end temperature of compression. This partition forms a 
part of the cover to the cylinder head proper, and this construction makes it com¬ 
paratively easy to get at the interior of the cylinder without removing the piston. 
Above 70 H.P. the pistons are fined with white metal, and above 100 H.P. the 
pistons are water-cooled. All of the sizes, with the exception of the small models 
up to 8 H.P., have an outboard bearing. Below 12 H.P. the electric ignition is 
replaced by the open hot tube. 

The double-acting 2-cxjde engine, Fig. 493, as stated, is not made below 350 H.P. 
Its method of operation is in general as follows .(see Figs. 494 and 495): 

The main cylinder a contains a water-cooled piston b the length of which is 
about | that of the stroke. A .side crank, 110° ahead of the main crank, by means 
of common gearing operates an air pump c and a gas pump d, which are controlled 
by piston valves c' and d' respectively. From each pump separate ports lead to 




















STATIONARY ENGINES 


335 


the inlet valves e and e', in the housing of which the mixture is formed. The ring of 
ports at the middle of the cylinder serves as the exhaust outlet for both ends of the 
cylinder, alternately on one and then on the other side of the piston b. At the 
moment that the piston commences to cover the exhaust ports, the pump pistons 
have already completed about half of their stroke. Up to this instant, the suction 
passage of the gas pump d has been kept open, thus allowing about 40% (at full 
load) of the gas to flow back into the suction mains. After this the valve d' 
closes the suction passage and opens the discharge passage, and the main cylinder a 
commences to take its charge. In the mean time the air pump c has delivered air 
into its discharge passages, or into a receiver in the base of the machine, from the 
beginning of its stroke and has forced some of this air up into the gas passage next 



Fig. 486. 


to the inlet valve housing e. Therefore, when the inlet valve e is opened, air will 
first flow out of both passages into the cylinder to be soon afterward followed by the 
mixture formed in the housing. This scheme gives a preliminary charge consisting 

of air only, which is desirable on several accounts. Scavenging and charging con¬ 
tinue until the pump pistons have reached their dead centers, which happens at the 
instant the piston b has covered the exhaust ports /. The mixture is next com¬ 

pressed and then ignited in two places simultaneously. 

The principal features of the construction are shown in Plates XII to XV. 

Speed regulation is effected by proportioning the quantity of gas to the load by 
means of changing the effective delivery stroke of the gas pump. For this purpose 
the hollow gas pump piston valve a, Figs. 496 and 497, contains a two-part piston 
valve d ' which controls the beginning of the delivery period in the stroke of the 
pump. The main valves a and b of the gas and air pumps respectively are carried 

by the common hollow valve rod c, which by means of the rocker arm c' is con¬ 

nected to the first eccentric. The auxiliary valve d of the gas pump is connected 
to the second eccentric by means of e and e f . 




330 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 

























































































































































































PLATE XII 



s.W'W't 


Water 


r 1 * " 



Erecting Plan for a Korting 2=cyc!e Blastfurnace Gas Engine. 

Rated H.P =500 at 100 r.p.m. 


Built by Gebr. Korting, Kortingsdorf bei Hannover. 


Atr 


Coofivq WaterSupph/fipe. ^ 

h- - 


- 1 - ^ 



7adffod used, 
only Z7z case of 
direct connection. 


/SZ<3 
























































































































































































































































































































































































































































































































































































































































































































PLATE XIII 



'///SA 


■'////////# 


///////////. 


V777777/' 




Length oyer A//-6J"3 


1/a/veG-ear 

7or B/o wina 

Cg/mder |jn^ 


. .: _ 1 — 



Section 

B-B 


Cas 


v 

Siegener Maschinenbau=Aktien=Gesellschaft vorm. 





Assembly Drawing and Erecting Plan, Porting 2-cycls 
Blast-furnace Gas Engine Direct-connected to Blowing Cylin¬ 
der. 


Rated H.P. = 500 at 100 r.p.m. 

Capacity of Blowing Cylinder is from 10 000 to 17 600 
cu.ft. per hour compressed to from 10 to 6 lbs per sq.in. 


A. and H. Oelchelhauser, in Siegen, Westf. 


































































































































































































































































































































































































































































































































































































































































































































































































































































































































































~ 






PLATE XIV 



Tr'.I” Inrinr-lf'/ 


Coofinq. Wafer Discharge 


HagrSft 


TICSEr 


|@®@ (§T '© @®©| 






fet 



Assembly Drawing, Korting 2=cycle Twin Engine, Rated Capacity 2000 H.P 

Built by the Maschinenbau-Aktien-Gesellschaft vorm. Gebr. Klein in Dahlbruch. 






































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































PLATE XV 



/fight 

Cghndtrrk 


_ _ Coo l 1 7 i<7 Wat&r-/nfot 

Compressed At 

Cooling y\fxter^ J ^ 


__//<b trtCas l/alve 

/^/ry I4x/ye ~5r&' ® 

Z^- 5 /on Coo/mg? WatsZ^> ^ 


Air Main. 


Built l>y the Siegener Maschinenbau-Akt. Gea, 
vorm. A. & H. Oechelhiiuser, Siegen. 


Inches. Stroke 55.1 Inches, 
g Blowing ('ylinder 69 Inches.) 


(Cylinder Diameter 31.5 
Diameter of Direct-connectin, 


(f l 





\V At 




L«j 




_ - — 

— - _ 























































































































































































































































































































































































































































































































STATIONARY ENGINES 


337 



ZCi 

% 


* 

^ O 

•N Tf 

.3 

b£) c 3 
G 


w 

CQ 

c 3 

C 5 


CM 

O 

© 

© 

GO 

o 


o 

o 


w 

ffl 

T 3 

0 > 

+- 

a 


r* ^ 

<r, 0 > 

£ CQ 

CQ S 
oj c 3 
^ £ 

2 0 

^.i 

T3 ~3 

o) 2 

3 g • 

S -g §> 

fcJD £ 5 

G -S G, 

-2 ^ w> 

S G 
O *p £ 
c n t 3 

.59 _, o 

S O 

c * * 

.2^3 

^ ^ N 

•S» 9 00 

s ^ 

Ph "2h bb 

oj rr 

J-. ^ 


































































































































































































































































































































































































































































338 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


The air-pump valve controls the suction and delivery of air independent of the 
load, making the alternate connections near the end of the pump stroke in each case. 
Thus the same quantity of air is handled at all loads. Gas valve a, on the other 
hand, keeps the suction passages of the gas pump open far into the stroke, so that of each 



Fig. 493. 


charge at least one-half is forced back into the suction mains. At maximum load, valves 
a and d close the suction passages at about half stroke, opening the delivery passages 
immediately thereafter. At this instant the main crank, 110° behind the pump crank, 
passes through its dead center. With decreasing loads, the governor delays the action 



of the auxiliary valve d, allowing the pump to force more gas back into the mains 
anc less into the cylinder. A modification of this scheme consists in allowing the pump to 
raw a part of the charge out of the pressure passages leading to the cylinder, thus 
controlling the quantity of fresh gas. The pressure in the pumps at the end of the 
delivery stroke goes up to 8.5 lbs. gauge. 

. ^ e _ ) a j r ^ orce< ^ U P i n t° the gas passages next the inlet valve, as above described, 
is intended first to act as a scavenging agent, and secondly to aid in “ stratifying ” 





























































































































STATIONARY ENGINES 


339 










































































































































































































































































































































































































I — r—g- 


340 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



I 2 

02 .O 

5 :“ 5 
lO Q 

o .. 


£ fc.' 

.g 

Q O 


e* as 

.r i 

t a 
3 « 

■Jh 

bO 













STATIONARY ENGINES 


341 



the charge in the cylinder. There is no doubt that by this means a comparatively 
lean mixture may be obtained next to the piston.* 


Table 51 



Figs. 499-501. —Erecting Plan for Standard 
Korting Pressure Producer Gas Installations. 

(In case anthracite is used the space E 
occupied by the sawdust purifiers is not 
required.) For suction-gas plants, the di¬ 
mensions C, D, and E may be reduced by 
from 4 to 


DIMENSIONS OF KORTING PRESSURE GAS INSTALLATIONS 


Dimensions in feet. 


A 
B 
\ C 

D 

E 


An (in B.H.P.). 

12 

14 

16 

20 

25 

30 

35 

40 

50 

60 

80 

Number of producers. 

1 

1 

1 

1 

1 

1 

1 

2 

2 

3 

3 


Single-unit Plants 


A 

15.6 

16.4 

17.0 

17.7 

18.4 

20.3 

21 .6 

22.0 

23.0 

24.2 

26.2 


B 

11.2 

11.5 

11.5 

12.5 

12.5 

12.8 

13.1 

14.1 

14.7 

15.7 

16.4 

Dimensions in feet. . • 

c 

10.5 

10.5 

10.5 

10.5 

10.5 

10.5 

10.5 

11 .8 

11.8 

13.1 

13.1 


D 

17.4 

17.4 

17.4 

17.4 

17.4 

17.4 

17.4 

18.3 

18.3 

19.7 

19.7 


E 

7.9 

7.9 

7.9 

7.9 

7.9 

7.9 

7.9 

8.8 

8.8 

9.8 

9.8 


Two-unit Plants 


15.6 

22.4 

14.8 

13.1 

9.8 


16.4 

23.0 

14.8 

13.1 

9.8 


17.0 

23.0 

14.8 

13.1 

9.8 


17.7 
25.0 

14.8 
13.1 

9.8 


18.4 

25.0 

14.8 

13.1 

9.8 


20.3 

25.6 

14.8 

13.1 

9.8 


21.6 
26.2 
14.8 
13.1 
9.8 


22.0 

28.2 

16.4 
15.7 

10.5 


23.0 

29.4 

16.4 
15.7 

10.5 


24.2 

31.4 

18.0 

18.0 

11.8 


26.2 
32.8 
18.0 
18.0 
11 .& 


* See the extended discussion by Reinhardt concerning charging and governing phenomena in 
Stahl und Eisen, 1902, No 21. 










































































































































































342 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Figs. 502 and 503.—Erecting Plan for a Korting Gasoline Engine, Model M-A, with Spraying Carburetor. 

Koom for the storage tank a and the feed tank c is separated from the engine room by a heavy wall. 

Tank c is filled by means of pump b.) 

Operating Results, (a) Illuminating Gas Engine, 8 B.H.P., Z) = 6.90", £ = 13.4" 
compression ratio £=6.41. Magneto Ignition, speed regulation by varying charge 
volume. Tests by Prof. E. Meyer.^ 

The illuminating gas used had a heating value of 496 B.T.U. per cu.ft., at 66.2° 
F. and 751 mm. bar., the consumption figures of the table, however, have been 
recomputed to the basis of gas of 562 B.T.U./cu.ft., at 32° F. and 760 mm. 


Table 52 


Test No. 

Load. 

R.p.m. 

n 

B.H.P. 

* r e 

Gas per 
B.H.P. 
Hour, 
cu.ft. 

Com¬ 
pression 
Pressure, 
lbs. per 
sq.in. 

Pressure 
at End of 
Suction, 
lbs. per 
sq.in. 

Ratio 
Charge to 
Cylinder 
Volume, 

%. 

Ratio Air 
to Gas. 

Economic 
Efficiency, 
% ' 
Vw 

Heat in 
Cooling 
Water, 

%. 

1 

2 

3 

full 

b 

no load 

221.2 
226.2 
229.3 

10.27 

5.25 

0 

16.05 
20.60 
57.2 
per hour 

139 

85 

35.5 

13,8abs. 
9.80 
6.35 

76 

50 

23 

6.76 

6.99 

6.00 

28 

22 

34.9 

38.6 


The diagram 
pressure of 85 lbs. 
lower loop diagram 
85-2.3=82.7 lbs. 


obtained at maximum load showed an average mean effective 
Per_sq.in. From this Meyer subtracts 1 2 the mean pressure of the 
Pi-- 2.3 lbs. per sq.in., and from the net pressure Pi = p i+ —p._ = 
per sq.in., he computes the corresponding maximum I.ELP. = 11.55 


1 Z. d. V. D. I., 1900, p. 332. 

2 The correctness of this method of computation is just at present the subject of an extended 

controversy published in the Zeitschrift des Vereins Deutscher Ingenieure (1905, pp 324 331 and 517) 
and started originally by an article due to Professor Riedler. At the present X’t 7 • 

opinion., appear to holdla balance. The question is importance' VZZlZ atfng to X ^ 

parat.ve performances of 2- and 4-cycle engines. 1. is quite hkely that the issue will be settled by 
the code for the testing of gas engines now in preparation. ^ 












































































STATIONARY ENGINES 


343 


and the gas consumption per I.H.P.-hour = 14.4 cu.ft. Figs. 504 and 505 show 


two no-load diagrams obtained on Test No 
The author has taken the follow¬ 
ing additional data on illuminating 
gas engines from the catalogues and 
other papers of the firm. The tests 
were all made by independent ex¬ 
perts, and the figures for gas con¬ 
sumption have been referred to gas 
under standard conditions. 


1"=120 lbs. per sq.m. 



Fig. 504. 



Table 53 


100 B.H.P. Illuminating Gas Engine. 

Electric Station at Gottingen. 

Electric 

Station. 

No. 1. 

No. 2. 

Meissen. 

Load. 

full 


full 


full 

full 

BHP . 

114.1 

76.8 

115.0 

76.9 

97.20 

97.30 

Gas per BHP. hour, cubic feet. 

14.10 

16.38 

15.32 

17.75 

14.70 

14.05 

Tilpnnnmio pffipipnCV. . 

32 

27.5 

29.5 

25.0 

30.5 

32.3 

1 






(6) Two 125 B.H.P. pressure producer gas engines of the older type in the 
Electric Station at Erlangen gave the following results: 


Engine No. 

Mean load. 

Fuel consumed per K.W.-hr., boiler and producer, lbs 
or 

Gas coke and anthracite, per B.H.P. hour, lbs. 

Cooling water per B.H.P. hour, lbs. 

Mechanical efficiency (n=appr. 126 r.p.m.), % . ... 

From the above figures we may obtain the following additional data: 


Fuel per I.H.P. hour, lbs. 

Indicated thermal efficiency % (taking H u = 13,500 

B.T.U./lb.).. 

Economic efficiency of entire plant, %. 

The figures guaranteed by the makers for the above installation were: fuel con¬ 
sumption, 1.71 lbs. per K.W.-hour; mechanical efficiency, 80%; and cooling water 
consumption not to exceed 110 lbs. per B.H.P.-hour. The difference in speed between 
full load and no load was 2% of the rated speed; for sudden changes of load amount¬ 
ing to 25%, the variation in the speed amounted to not quite 1.5%. During the 
two-day tests each unit was temporarily loaded with 94 K.W. =appr. 140 B.H.P. 


162.5 

.746 

25.3 

21.4 


I II 

83.93 K.W. = 125 B.H.P. 88.7 K.W. = 136 B.H.P. 
1.230 1.29 

.837 .878 

73.6 72.5 

85 85 


































344 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


(c) A suction gas plant of 100 B.H.P. rated capacity, operated with brown coal 
briquettes tested in the summer of 1904 by R. Pawlikowski, showed the results given 
in the following table: 


Cylinder diameter, 21.2", 
Trial... 

stroke, 34.4", r.p.m., 150. 
. I 

II 

III 

Duration, hours. 

. 5 

3 

3 


. 150-132 

152-148 

156-154 

B.H.P. 

. 118.5-113.5 

108.5 

53.8-53.6 

Suction vacuum, inches water. 

. 1.97-2.36 

1.57-1.93 

.97-1.38 

Temperature of cooling water, outlet, degrees F. 

. 104-113 

95-100 

91-95 

Weight of one full charge for producer, lbs. 

. 154-198 

154-176 

88 

Time of gasification per charge, hour . 

. 1 

1 

1 

Fuel consumption per B.H.P. hour, lbs. 

. 1.477 

1.492 

2.18 

Fuel consumption per B.H.P. hour, lbs., guaranteed .. 

. 1.562 

1.562 

2.39 


At one-quarter load the fuel consumption was guaranteed not to exceed 3.19 lbs. per 
B.H.P.-hr. During stand-by periods lasting for 12 hours (over night) the producer 
burned from 99 to 110 lbs. of coal, which amounts to about 8% of the daily con¬ 
sumption at full load. 

Kind of fuel: Brown coal nut briquettes, average size 2", heating value 8607 
B.T.U./lb. 

3. Vereinigte Maschinenfabrik Augsburg und Maschinenbau-Ges. Niirn- 
berg A.-G., Werk Niirnberg. (See Plates XVI to XVIII.) 

The new model of large gas engine made by this firm and recently introduced 
operates on the 4-cycle double-acting principle and uses eccentrics to operate the 
valves. The constructive details of these engines are clearly brought out by plates 
following, and by Figs. 508 and 509. The sizes at present built and their main dimen¬ 
sions are compiled in the following table. 

Table 54 

STANDARD CAPACITIES AND MAIN DIMENSIONS OF NEW TYPE DOUBLE-ACTING LARGE 

GAS ENGINES 


Single-cylinder units. 

D. 6 a 

D. 7 

D. 7 a 

D. 8 

D. 9 

D. 10 

D. 11 

D. 12 

D. 13 

Rated B.H.P. 


160 

220 

245 

315 

400 

480 

550 

690 

845 

R.p.m. 


150 

150 

125 

125 

120 

110 

110 

94 

90 

Max. length to center of crank- 










shaft.. 

.ft. 

18.0 

19.2 

20.3 

22.0 

23.8 

25.6 

28.5 

30.2 

32.8 

Diameter of fly-wheel. 


13.5 

16.1 

16.1 

16.7 

16.7 

18.1 

20.0 

21.3 

22.1 

Distance floor to center 

of 










shaft. 

.ft. 

2.46 

2.51 

2.54 

2.62 

2.79 

2.95 

3.12 

3.20 

3.28 

Maximum width*. 

.ft. 

14.7 

15.4 

18.1 

19.7 

20.5 

21.4 

22.1 

23.0 

23.8 

Tandem units. 

D. T. 6o 

D. T. 7 

D. T. 7a 

D. T. 8 

D.T.9 

D.T. 10 

D.T. 11 

D.T. 12 

D.T. 13 

Rated B.H.P. 


350 

480 

530 

685 

870 

1050 

1200 

1500 

1850 



150 

150 

125 

125 

120 

110 

100 

94 

90 

Max. length to center of crank- 










shaft. 


28.0 

30.3 

32.8 

35.2 

38.2 

41.0 

44.3 

46.7 

49.2 

Diameter of fly-wheel. 

.ft. 

13.5 

16.1 

16.1 

16.1 

16.7 

18.1 

20.0 

21.3 

22.1 

Distance floor to center 

of 










shaft. 


2.46 

2.51 

2.54 

2.62 

2.79 

2.95 

3.12 

3.20 

3.28 

Maximum width*. 

.ft. 

18.0 

18.7 

19.7 

20.5 

21.3 

22.1 

23.0 

23.8 

24.6 


* To outer edge of out-board bearing. 















































STATIONARY ENGINES 


345 


The sketches shown in Figs. 506 and 507 indicate the manner of opening up the 
interior of the cylinder and the taking out of the pistons in a Niirnberg tandem 
engine. 


Fig. 506. —Method of Opening 
up the Front End Combustion 
Chambers. 



Fig. 507. —Taking the Pistons 
out of a Tandem Engine. 




rZSOt 


-IS 


-S 


Fig. 510. 


Operating Results. Recent information on fuel consumption and capacity obtained 
by independent observers is lacking in the data available. The heat consumption at 
full load for the large machines is given at from 
8700 to 10 300 B.T U. per B.H.P.-hour. This, if a 
gas producer is used, would mean a consumption of 
from .815 to .945 lbs. of anthracite and from 1.00 
to 1.17 lbs. of coke per B.H.P.-hour. The consump¬ 
tion figures for f and ^ loads are stated to be 10 
and 30% higher respectively than at full load, which 
would be a very remarkable showing. 1 The diagram, 

Fig. 510, was taken from a blast furnace gas engine. 

4. Deutsche Kraftgas-Gesellschaft m.b.H. in Berlin, (see also Plates XIX 

to XXH). , . + 

This company are the owners of the Oechelhauser 2-cycle engine patents. These 

engines are at present built in single cylinder sizes from 250 to 1000 B H.P. and as 
twin units from 500 to 2000 B.H.P. The speed of all of the sizes is 125 r.p.m. ih « 
fuel is mostly blast-furnace gas. The 1000 B.H.P. twin cylinder engine for a coefficient 
of regulation of (operation of alternating current generators m P aia e 

fly-wheel weighing only 28 tons. The indicator card, when operating with blast-furnace 
gas, shows a compression pressure of about 150 lbs. and at maximum oa an exp osion 


i Since the printing of the above statement, Professor Riedler has published in the Zeitschrift d. 
V D MflMDOTl artide on “Large Gas Engines/’ from which the writer, among other things, 
takes the flowing results obtained in the operation of a double-acting tandem Numberg engme (No. 
IV, D = 33.5", 5 = 43.4", n=106): 


Ne= 276 

pi= 60.8 

rjm= 48.5 
Consumption = 18 800 


550 860 

42.5 60.1 

69.0 76.1 

12 400 10 920 


At the low T est load (276 B.H.P.) one 


1024 
68.1 
79.0 
10 120 

cylinder was 


1100 
70.4 
82.1 
9780 
cut out. 


1131 1170 B.H.P. 

73.2 75.2lbs. per sq.in. 

82.6 83.1% 

9320 9080 B.T.U. per B.H.P. per hr. 

The power developed was measured 


electrically. 





























































346 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Fig. 508—Niimberg Tandem Engine direct connected to Blowing Cylinder. Effective H.P.=870 at 120 r.p.m. (D=33.5", S=43.4".) 

View of valve-gear side. 


















STATIONARY ENGINES 


347 




xn 

*> 

v 

O 

c6 

co 



rH 

£ 

4* 


O 

CM 


+3 

c 3 

O 

I- 

CO 


hH 

a 


> 


w 


OJ 

^3 


>> 

o 

bX) 

q 

£ 

3 

o 


~3 

<D 

-4^> 

o 

o 

G 

G 

O 

o 

i 

o 

- 


0) 

a 

• r-H 

bX) 

G 

H 


<u 

£3 

G 

c3 


H 


b£> 

a> 

-2 

- 

:P 


1 


o 

o 

o 

6 

•-< 


View showing Frame Construction, Blowing Cylinder, etc. 
















348 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 






















































































































































































































































STATIONARY ENGINES 


349 



illustrated on Plates XVI and XVII. 








































































































































































































































































350 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 








































































































































































PLATE XVI 



Assembly Drawing, 2000 H.P. Double=acting Tandem 4=cycle Engine. 


Vereinigte Maschinenfabrik Augsburg und Maschinenbaugesellschaft Niirnberg A.G. Werk, X urn berg. 

(See also Plate XVII.) 





































































































































































































































































































































































































































































































































































































































































































































































































PLATE XVII 



Assembly Drawing, 2000 H.P. Double=acting Tandem 4=cycle Engine. 

Vereinigte Maschinenfabrik Augsburg und Maschinenbaugesellschaft Niirnberg A.-G. Werk, Niirnberg. 
(See Elevation and Cross-section, Plate XVI. For details of regulation see Fig. 326.) 


vvfjr 

.iprrj 

pi 




r 

(Q - 

• 

. . . . 


@ 


• • 0 • 






@Y5t 



V-l— 

s 


pr 

©J 

§> 































































































































































































































































































































































































































































































































































































































PLATE XVIII 



I Compress e ct 
Air 


Erecting and Piping Plan for one of the older 
Types of 4=cycle BIast=furnace Gas Engine. 

Built by the Vereinigte Maschinenfabrik Augsburg and 
Maschinenbau-Gesellschaft Xurnberg 
A.-G. Werk, Niirnberg. 

Rated Brake Horse-power 700 at 90 r.p.m. 

(Cylinder Diameter 52 Inches. Stroke 55 Inches.) 


For Assembly Drawing of Engine and further Information 
see page 350. 


























































































































































































































































































































































































































































/./<5 
















- r 


~~i — 



/'- 





-\\ 


fH 

@ 

. 

© 



Vi 


© 

— 


j © 

4 


T irgr~Ts~ r^tr ts ii - _ ii*?, 7 

©II I . I : ! "■! © n-liy_ _^U_LlA ' 

! _ •} 

if 'll ■ 

ij _ r' =L k ^ft- _i-»; _ 

7 I I Bbmp 

1 1 I ' l i j | Be/o^v /^Zoor- 

- -n 


(Vi'* 1 -^ / 

II rcv/l 


Rated horse - power = 500 at 125 
r.p.m. , _ 

For details of same engine driving 
blowing cylinder, see Plate XX. 


II 1 

Ti««*-!:&- f - 


Gas. 


i Ml 


- il t£v Tp, i f,T, f ■«»:; - Azr 

©|l _ I,-, ? 7, 1 _ S') tJ -—Tw=;-i 

V_ _ _____ Jk-J, 

-7 — 


rupn 


'jft: 

! ' 


Ol 


. U—^ 


. _ J_I 


* n .T', 




Concern mg Speed Regulation and 
other details, see Plates XXI and 
XXII. 




Assembly Drawing, Oechelhiiuser 2=cycle Blast=furnace Gas Engine. 



































































































































































































































































































































































































































































































































































































































































































































































































































































































T/.S- 



PLATE XX 


Starting Va/ve 




■ 

p * ■' =■ i < i ■ 



‘ 




- \ 




---- - 




Rated Horse-power = 500 at 100 


For details of same Engine driving Electric 
Generator, see Plate XIX. 


53552 


w 

!— () 



5 T 




f== 

—0— 

IS 

© 

0 


^-»■■■■■! 'Mf 



# 

Hi 

0 


123 






Assembly Drawing, Oechelhauser 2=cycle Blast=furnace Gas Engine 



































































































































































































































































































































































































































































































































































































































































































































































































































































PLATE XXI 


Leqend: 


a = Exhaust ports; 

6 = Air inlet ports; 

c = Gas inlet ports; 

6' = Cover ring for air inlet ports b\ may l>e adjusted 
by hand through linkage d, to regulate the 
mixture; 

d = Cover ring for gas inlet ports c, controlled by 
the governor through the linkage e and e'. 
The cutting off of the ports commences on 
the side of the cylinder opposite the igniter 
in order to retain a rich mixture around the 
latter; 


Under control of the gov¬ 
ernor at the beginning 

of the discharge strokes 
/= Gas-return valve; , , , 

of the pumps in order to 
/' = Air-return valve; , , 

decrease the pump work 

with a decrease in the 

load on the engine. 

g = Starting valve, operated from the igniter gear; 
hh' = Make-and-break igniters; 

i = Control ring for air receiver, operated by eccen¬ 
tric through %'. The purpose of this is to 
separate the scavenging air as far as possible 
from the air used for the mixture, in order 
to prevent the latter from getting too lean. 


Assembly Drawing, 1500 B.H.P. Oechelhauser 2=cycle Engine. Built by A. Borsig, BerIin=Tegel 






















































































































































































































































































































































































































































































































































































































































































PLATE XXII 



mm\\ 




* ® 

■v- 

® * 

® • • ® 

A 

Vi 

*| 


L . 71 


r ^—>1 





E-- 

-?-7 


// / 
V/ 


■Ol) 

r*—- ' 

Jp 




',-ru ‘ , ' 


Legend: 

a = Exhaust ports; 

t> = Air inlet ports; 

c = Gas inlet ports; 

6' = Cover ring for air inlet ports b, may he adjusted 
by hand through linkage d, to regulate the 
mixture ; 

c'=Cover ring for gas inlet ports c, controlled by 
the governor through the linkage e and e'. 
The cutting off of the ports commences on 
the side of the cylinder opposite the igniter 
in order to retain a rich mixture around the 
latter; 


l nder control of the gov¬ 
ernor at the beginning of 

/= Gas-return valve; the dischar g e st ™ k * s of 

/' = Air-return valve; the PUmpS “ order to de ‘ 

crease the pump work 

with a decrease in the load 
on the engine. 

g = Starting valve, operated from the igniter gear; 
hh = Make-and-break igniters; 

i = Control ring for air receiver, operated by eccen¬ 
tric through i'. The purpose of this is to 
separate the scavenging air as far as possible 
from the air used for the mixture, in order 
to prevent the latter from getting too lean. 


Assembly Drawing, 1500 B.H.P. Oechelhauser 2=cycle Engine. Built by A. Borsig, Berlin=Tegel. 























































































































































































































































































































































































































































































































































































































STATIONARY ENGINES 351 

pressure of about 300 lbs. per sq.in. For method of speed regulation, see Plates XXI 

and XXII. 

The method on which this engine operates is best explained from Plate XXII. The 
cylinder, a long straight tube open on both ends, contains two pistons. The front piston, 
by means of rod, cross-head, and connecting-rod, is connected to the middle crank of 
the three-throw shaft. The back piston, by means of a yoke, side rods, etc., is connected 
to the two outside cranks. This construction results in an engine almost perfectly 
balanced. The front piston controls the exhaust ports <z, while the back piston uncovers 
first the scavenging air ports b, and a little later also the gas or mixture ports c. The 
air and gas pumps are in this case placed at the side of the engine and operated 
by a rocker arm and shaft, as shown. In some of the other constructions these pumps 
are operated by an extension of the back piston rod, or they may be placed under the 
floor and operated by rocker arm. The pumps supply gas and air under certain pressure 
to individual receivers which in turn connect with the gas and air ports c and b re¬ 
spectively. To explain the operation, suppose that both pistons have completed their 
instroke and have compressed between them the fresh charge (180° from the crank 
position shown in the plate). The charge is ignited electrically at g and the explosion 
forces the pistons apart. Near the end of the stroke, the front piston first uncovers 
ports a, and exhaust commences. A moment later the back piston uncovers ports b, 
admitting a charge of fresh air to sweep out the cylinder toward the front. This is 
next followed by gas from ports c to form the mixture. On the return the ports are 
covered in succession and the operation is repeated. 

Operating Results. A 500-H.P. Oechelhauser blowing engine in the Borsig Works 
was tested by Prof. E Meyer. The fuel was coke oven gas. 

The machine was originally built by A. Borsig-Tegel for operation with blast¬ 
furnace gas. Hence to operate it with coke oven gas called for certain changes in 
pumps, ports, regulation, etc., w r hich changes left certain imperfections in the construc¬ 
tion, at times affecting the results unfavorably. 

The main tests were made during the forepart of August. Certain check tests on 
the most important of the main trials were made at the end of October, 1904.* 


Power cylinder 


MAIN DIMENSIONS OF ENGINE 


f Diameter. 26.6" 

■) Stroke front piston . . . 37.49" 
l Stroke back piston . . . 37.31" 


Air pump i 


l Diameter . . . 44*. 85" 
1 Stroke.19.7" 


, f Diameter 
Blowing cylinder | gtroke 


65.0" 

37.31" 


/ Diameter ... 23.20" 
Gas pump j g troke . 19 . 7 " 


ANALYSIS OF THE COKE OVEN GAS USED (Vol. %) 


C0 2 . 4.91 

Heavy hydrocarbons.... 2.63 

0 2 . /. 20 

CO. 11.84 

H 2 . 42.00 

CH 4 . 19.73 

N. 18.69 


4.90 

5.30 

1.80 

2.10 

.30 

.40 

10.60 

10.20 

48.08 

43.80 

18.43 

20.30 

15.89 

17.90 



* The entire report has been published in Z. d. V. D. I., 1905, p. 324. 














352 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Fig. 518. Twin Two-cycle Engine, Oechelhauser Type. Rated B.H. P. = 1000 at 125 r.p.m., direct-connected to Alternating Current Generator. 

Built by the Ascherslebener Maschinenbau-Akt.-Ges., Aschersleben. 


























STATIONARY ENGINES 353 

The lower heating value of the gas was determined at short intervals of time by means 

of a calorimeter and was found to vary within the limits of 349 and 433 B T U 
per cubic foot. 


Table 55 


PRINCIPAL DATA FROM SEVERAL OF THE TESTS 


Tests Made 


August 4 , 5 , 1904 . 


Test Number. 


R.p.m.. 

Compression, lbs. per sq.in. 

Mean ind. pressure, lbs. per sq.in. 


P*+ . 

I.H.P. (from pi+), Ni+ . 

Mean ind. pressure in air / Front 
pump, lbs. per sq.in. . { Back 

Air pump, I.H.P., Ni . 

Mean ind. pressure in gas pump, 

lbs. per sq.in. 

Gas pump, I.H.P., N g . 

Total pump H.P., N +N g . 

Indicated engine H.P. = Ni 

Ni+-(Ni+N g )*. 

I.H.P. of blowing cylinder, Nw . 

^ . Nl + Ng, 

Ratl ° - N . . % 


Total eff.: i ? = 

Ni+ 

.% 

Mech. eff.: i)m =—. 

Nl 

. % 

Work of friction, Nr- 

= Ni—N . 

Heat used per I.H.P. 

hour: 

Based on Ni +.... 

. . . B.T.U. 

Based on Ni. 

. . .B.T.U. 

Based on Nw . 

. . .B.T.U. 


VII5 

XV 

XIV 

XIX 

XVII 

X 

IV 

VIII6 

XII 

110.6 

113.6 

112.9 

108.3 

96.4 

84.5 

86.8 

86.1 

68.4 

153 

150 

150 

143 

145 

143 

139 

139 

142 

74.5 

60.5 

57.1 

41.2 

32.6 

70.1 

53.4 

52.0 

57.1 

865 

723 

676 

468 

330 

622 

486 

470 

410 

7.3 

6.7 

6.8 

5.06 

5.85 

6.93 

4.70 

4.60 

4.10 

4.66 

4.36 

4.32 

3.44 

3.58 

3.86 

3.01 

3.09 

2.37 

104.0 

98.6 

98.7 

1 72.5 

71.5 

71.7 

52.6 

52.1 

34.8 

3.11 

3.45 

3.52 

3.56 

3.69 

3.53 

3.15 

3.34 

3.28 

7.25 

8.25 

8.37 

8.14 

7.51 

6.29 

5.76 

6.06 

4.73 

111.2 

106.8 

107.0 

80.6 

79.0 

78.0 

58.3 

58.6 

39.5 

754 

616 

569 

387 

251 

544 

428 

412 

370 

619 

476 

433 

269 

144 

457 

343 

338 

293 

14.7 

17.3 

18.8 

20.8 

31.4 

14.3 

13.7 

14.1 

10.6 

70.3 

65.7 

64.1 

57.5 

43.6 

73.5 

70.6 

71,8 

71.3 

82.1 

77.2 

76.1 

69.4 

57.3 

84.0 

80.2 

81.8 

78.9 

135 

140 

136 

118 

107 

87 

85 

74 

77 

6700 

6500 

6500 

7060 

9630 

7830 

6980 

6670 

6820 

7710 

7670 

7740 

8540 

12610 

8970 

7950 

7630 

7540 

9400 

1 

9900 

10140 

12270 

22100 

10610 

9900 

9310 

9580 


* See foot-note, p. 342. 


In the second series of tests, October, 1904, the best figure obtained for the heat 
consumption was 7310 B.T.U. per H.P. hour, based on N it and 8760 B.T.U., based on 
N w . The corresponding thermal efficiencies 34.8 and 29% respectively, are extra¬ 
ordinarily good. The quantity of cooling water used was 60.5 lbs. per H.P. of blowing 
cylinder, while the heat carried off by the cooling water amounted to 16% of the heat 
in the gas. These too are very excellent results. The quantity of cylinder oil used 
in the power cylinder was 1.19 lbs. (about .16 gallon) per hour. The other parts to be 
lubricated received per hour 2.72 lbs. (about .37 gallon) of fresh oil besides a certain 
quantity of recovered oil. 

5. Louis Soest & Co. in Reisholz-Diisseldorf. (Plate XXIII.) 

As a result of its excellent constructive features (compare the main parts shown in 
Figs. 87-89, 144-147, and 276 in Part II) this new engine has made a name for itself 
in very short time. One of the two 600-H.P. twin engines shown on Plate XXIII 
was exhibited in Diisseldorf in 1903. The two-throw crank-shaft, forged from one solid 
block, weighed 32100 lbs., of which 5700 lbs. belonged to the balance weights. The 




















































354 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


25af 


diameter of the wheel, for a coefficient of regulation of y^th was 19.7 ft. (F ~125 ft. 
per sec.). The weight of the wheel was 66000 lbs., corresponding to 110 lbs. per rated 
horse-power. The indicator card, Fig. 519, was obtained when operating on a blast¬ 
furnace gas carrying about .0006 gram of dust per cubic foot and having a heating 
value of 95.5 B.T.U. per cu.ft. The mean effective pressure of this 
diagram was about 71 lbs. per sq.in., corresponding to a horse-power 
developed of 660 B.H.P. 

Late in 1903 this firm took up the manu¬ 
facture of double-acting 4-cycle engines for the 
largest sizes and is at present engaged in 
developing a double-acting 2-cycle machine 
(similar to the Korting, but having the firm's 
own method of forming the mixture and of 
governing). A 400-H.P. machine of this type 



Fig. 519. 


is already being used for experimental work. 

6. Giildner-Motoren-Gesellschaft in Aschaffenburg 

XXV). 


(See Plates XXIV and 



The degree to which the efficiency and specific capacity of a given type of internal- 
combustion engine may be raised by intelligent treatment of the constructive features 
is well shown in the new Giildner engine. Fundamentally nothing but a simple 4-cycle 
engine, this machine owes the unusual efficiency 
and capacity shown in the tables below entirely 
to careful attention to all its details, and 
especially to the design of the compression and 
combustion chamber, and the mixing, ignition, 
and governing arrangements. Fig. 520, together 
with Plate XXIV, show the main construc¬ 
tive features of one of the 100-H.P. engines. 

Probably the most striking feature is the 
double wall A-frame, which is a development 
of the author’s own older designs for the first 
Diesel engine, operating without a cross-head 
and which of late has become well known in 
the newer Augsburg design of this type of 
machine. 

For reasons strongly emphasized in all parts 
of this book, the Giildner engine at all times 
operates with a mixture as pure and as uniform 
as possible and in every respect brought to the 
best condition attainable for most efficient com¬ 
bustion. One means to this end is the tem¬ 
porary combined action of the exhaust and 
the inlet valve such that the combustion chamber 
is filled with air even before the suction 
stroke proper is commenced. It follows that the cylinder will get a greater quantity of 
cool and pure mixture, which, at all loads on the engine, will ignite with certainty and 
burn efficiently. (See the developed regulation diagram, Fig. 521.) The speed of the 
engine is regulated by means of the governor and gear shown in Figs. 327 and 328, 
p. 245. This arrangement operates to proportion the quantity of charge to the load; 


Fig. 520.—100 H.P. Giildner Engine. 
(Z> = 18.7", 5 = 27.5", « = 160.) 










PLATE XXIII 




Plan of a 1200 H.P. Blast=furnace Gas Installation. 

(Two Twin 4-cycle Engine, 600 B.H.P. each. Cylinder Diameter 36", 

Stroke 42".) 

Built by Louis Soest & Co., Reisholz, DQsseldorf. 

For details see Figs. 87-89, 144-147, and p. 353. 







































































































































































































































































































































































































































































































































































































































PLATE XXIV 




Key to Main Parts: 
a = Inlet, mixing, and governing valves; 
b = Exhaust valve (water cooled); 
c= Electric igniter; 
d = Magneto; 

a r i,' } c ! , and d', respective operating levers; 


Assembly Drawing, 100 H.P. Giildner Suction Gas Engine. 

Built by the Giildner Motoren-Gesellschaft Aschaffenburg. 



Ivky to Main Parts: 

e = Starting valve (hand operated); 

/= Throttle valve to regulate cooling 
water to exhaust valve; 

g and g’ = Governor (for details see pages 327 
ai.d 328); 

A = Oil pump for piston and piston pin. 


























































































































































































































































































































































































































































































































































































































































































































PLATE XXV 



rrr rr-f 


~F. <?.. s.; * T T- r , " ; < r 


300 H.P, 


130 H P 


d 


Til 

T 



T 


1 








■ %ZK7////& 


7%j/ ///////)7 / 


'flZZZZ&ZZZL 


jjii ii i 




JKZ7A 


Patels 


-1. 



jeH/=> 

Oaso/iTte 




Plan of a Guldner Suction Gas Plant. Total Capacity 1500 H.P. 

(4 Engines 300 H P. each, 2 Engines 150 H.P. each; 1-12 H.P. Auxiliary Engine for Compressors, Pumps, etc.) 

The general features of this plan, especially the length of the building, were dictated by the shape of the available plot. 





















































































































































































































































































































































































































































































































































































































































































































































































































PT.ATrc YWT 



Assembly Drawing and Erecting Plan, Diesel Oil Engine. 

Rated H.P. = 70 at 1G0 r.p.m. 

Built by the Vereinigte Maschinenfabrik Augsburg und Maschinenbaugeseb.schaft Nurnberg A.-G., Werk Augsburg. 





















































































































































































































































































































































































































































































































































































































































STATIONARY ENGINES 


355 


but to obviate as far as possible the difficulty encountered in igniting the mixtures at 
the lower loads, brought about by lower compression pressures, the governor also 
controls the proportion of gas in the charge. The success of this system of governing 
explains why the heat consumption per horse-power-hour of these engines shows such a 
comparatively slow increase with decrease of load. 

The erecting plan, Figs. 522 and 523, shows the practical disposition of the apparatus 
making up a suction-gas plant, arranged with a view to accessibility in all its parts. 
Attention is also called to the short lengths and accessibility throughout of the gas 
main, important points in themselves. In the larger installations the exhaust pipe is 
jacketed and the jacket is used directly to conduct away the cylinder cooling 
water. 

Whenever possible the exhaust muffler is placed near the roof either in the engine 
or in the producer room, the idea being to promote ventilation by thus heating the 
upper layers of the air. For the same purpose, the inlet of the suction pipe to the 



engine is placed in the producer room. (For details of] Guldner producers, see 

p. 280.) 

The gas and air controlling valves of the large Guldner machines may be operated 
from the floor as well as from the platform. In these sizes the exhaust valve is also 
water cooled. All of the sizes made have ignition gear that may be set by hand, 
pressure oil pumps for the pistons, and a central'oiling system for the bearings. Up 
to 20 H.P. the starting is accomplished by means of the starting gear shown in 
Figs. 443 and 444, p. 291; the larger machines are started by compressed air obtained 
with the small compressors illustrated in Figs. 451-453, p. 294. The details of some of 
the other machine parts are shown in some of the type constructions illustrated in 

Part II. 

At present Guldner engines are built as single cylinder engines with capacities 
ranging from 10 to 125 H.P., and two cylinder units are made in sizes from 100 to 
250 H.P. The main dimensions of the engines as well as the room required under 
ordinary conditions, are shown in Table 56 following. 









356 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


Table 56 

PRINCIPAL DIMENSIONS OF SINGLE-CYLINDER GULDNER SUCTION GAS PLANTS 


(See Plans, Figs. 522, 523) 


Nn in B.H.P. . 

10 

12 

15 

20 

25 

30 

35 

40 

50 

60 

70 

80 

100 

125. 

R.p.m. 

240 

240 

225 

225 

210 

210 

200 

200 

190 

180 

170 

160 

155 

150 


A 

7 

87 

8 

20 

9 

.20 

9 

85 

9 

85 

10 

.5 

10 

.5 

11 

.80 

12 

.75 

13 

.75 

14 

.75 

15 

70 

18 

.0 

21 

.30 

U 

o 

B 

3 

28 

3 

40 

4 

.28 

4 

60 

4 

.93 

5 

25 

5 

.91 

6 

.33 

7 

.07 

7 

.87 

8 

55 

9 

18 

9 

84 

12 

10 

o 

p 

C 

4 

60 

4 

93 

5 

58 

5 

91 

5 

91 

6 

23 

6 

57 

6 

90 

7 

22 

7 

.57 

8 

20 

8 

53 

9 

52 

10 

50 

-a 

P 

D * 

10 

50 

10 

50 

11 

80 

13 

10 

13 

75 

15 

10 

16 

40 

16 

40 

16 

40 

17 

40 

19 

CO 

20 

35 

21 

30 

22 

30 

X «+-. 

E 

6 

92 

7 

20 

7 

70 

8 

05 

8 

53 

8 

85 

9 

85 

9 

85 

9 

85 

10 

50 

11 

50 

12 

50 

13 

10 

14 

40 


F 

1 

13 

1 

65 

2 

62 

2 

78 

2 

78 

3 

45 

3 

12 

4 

43 

5 

08 

6 

08 

5 

25 

6 

23 

6 

90 

8 

85 

• P o 

G 

6 

55 

6 

55 

6 

55 

7 

05 

7 

05 

7 

05 

7 

35 

7 

35 

7 

70 

7 

70 

9 

50 

9 

50 

11 

15 

11 

50 

fcJD Sm 

S 3 

H 

5 

29 

5 

29 

5 

29 

6 

27 

6 

27 

6 

27 

7 

06 

7 

06 

8 

05 

8 

05 

9 

85 

9 

85 

11 

50 

13 

75 

" b* 

I 

1 

97 

1 

97 

1 

.97 

1 

97 

1 

97 

1 

97 

1 

97 

1 

97 

1 

97 

1 

97 

1 

97 

1 

97 

2 

62 

2 

62 

O 

K 

2 

62 

2 

62 

2 

.62 

2 

96 

2 

96 

2 

96 

3 

28 

3 

28 

3 

61 

3 

61 

3 

94 

3 

94 

4 

27 

4 

58 

II 

L 

1 

64 

1 

64 

1 

.64 

1 

64 

1 

64 

1 

64 

1 

64 

1 

64 

1 

64 

1 

64 

1 

64 

1 

64 

1 

97 

1 

97 

*1 2 

M 

5 

.08 

5 

24 

5 

58 

5 

58 

5 

90 

5 

90 

6 

25 

6 

25 

6 

55 

6 

88 

7 

55 

8 

20 

9 

50 

11 

15 

<D T3 

N 

3 

.22 

3 

28 

3 

.44 

3 

72 

3 

77 

3 

94 

4 

03 

4 

27 

4 

59 

5 

08 

5 

37 

5 

63 

6 

07 

6 

55 

.1 3 

0 

5 

.57 

5 

90 

6 

.56 

7 

38 

7 

87 

8 

53 

9 

50 

9 

50 

9 

85 

10 

80 

11 

15 

11 

80 

11 

80 

12 

15 

£ 

P 


.36 


.38 


.39 


39 


41 


43 


44 


46 


73 


79 


86 


92 


98 

1 

05 

3 

Q 

1 

.97 

1 

.97 

2 

.21 

2 

46 

2 

95 

3 

28 

3 

62 

3 

77 

3 

94 

4 

27 

4 

58 

5 

24 

5 

73 

6 

23 

§ 

R 


.92 


.92 

1 

.07 

1 

15 

1 

31 

1 

37 

1 

44 

1 

51 

1 

80 

2 

10 

2 

39 

2 

52 

2 

95 

3 

28 

Width of 

belt 


.43 


.43 


.46 


53 


.59 


62 


66 


69 


82 


95 

1 

08 

1 

15 

1 

34 

1 

51 


* This amount of head-room is ample to allow of the taking out of the piston together with the 
connecting-rod by means of a crane. Where this room is not available the piston may be taken out by 
itself, in which case the head-room may be made equal to about f D. 


Operating Results. (a) The first vertical Guldner engine in actual operation was 
tested in 1903 by Prof. M. Schroter with the assistance of Dr. Koob. The tests were 
conducted with both illuminating and anthracite suction gas. The engine was rated at 
20 H.P., cylinder diameter = 9.87", stroke = 15.75". The maximum compression was 113.5 lbs. 
with illuminating gas and 157 lbs. with suction gas. In both cases the indicator 
diagrams showed a mean effective pressure in excess of, 113 lbs. per square inch at 
maximum load. The exhaust gases at all loads were invisible and odorless. 

Table 57 


RESULTS OF TESTS WITH ILLUMINATING GAS 


Approximate Load. 

Meaii 

R.p.m. 

Mean 
Effective 
Pressure, 
lbs. per 
sq.in. 

Average 

I.H.P. 

Lower 
Heating 
Value of 
Gas, 
B.T.U. 
pei- cu.ft. 

Consumption of 
Gas Referred to 
32° F. and 
735.5 mm. Hg. 

Consumption of 
Gas Reduced to a 
Heating Value of 
560 B.T.U. 
per cu.ft. 

Heat 
used per 
I.H.P. 
Hour, 
B.T.U. 

Indi¬ 

cated 

Thermal 

Effici¬ 

ency, 

J;i, 

% 

Per 

Hour, 

cu.ft. 

Per 

I.H.P. 

Hour, 

cu.ft. 

Per 

Hour. 

Per 

I.H.P. 

Hour. 

Half maximum load ... 

211.8 

59.5 

19.1 

488 

299 

15.60 

275 

14.40 

8640 

31.6 

Same. 

213.9 

63.7 

21.7 

496 

295 

14.25 

279 

13.45 

7550 

33.9 

\ maximum load. 

212.8 

95.3 

30.9 

495 

384 

12.43 

362 

11.72 

6580 

38.8 

Same. 

213.7 

94.0 

30.6 

501 

383 

12.38 

360 

11.80 

6620 

38.6 

Nearly maximum load . 

214.5 

114.5 

37.1 

497 

457 

12.28 

436 

11.73 

6590 

39.0 

Same. 

210.7 

110.0 

35.3 

498 

399 

11.25 

377 

10.65 

5970 

42.7 


The heat balance for the last line of the above table reads as follows: Of the heat 
supplied in the illuminating gas, 42.7% are transformed into indicated work, 33.2% are 















































STATIONARY ENGINES 


357 



Figs. 522 and 523.—Erecting- and Floor-plan for Single-cylinder Guldner Suction-gas Plants. 
(See Table 5G for Dimensions.) 


lost in the cooling water, while 24.1% (remainder) are lost in exhaust and by 
radiation. 

The same high thermal efficiency is also shown when operating with suction gas. 
As Table 58 shows, the first machine to be operated on this gas showed an indicated 
thermal efficiency of nearly 29%, which includes all losses in producer and scrubber. 
This corresponds to a heat consumption of from 6750 to 7150 B.T.U. per I.H.P. per hour 
for the engine alone, which would, in round numbers, be equal to .55 lb. of anthracite 
per I.H.P., a figure which was afterwards found in actual service (see p. 359). 

In Table 58, “ Total” anthracite consumption includes coal used for starting, 
etc. “ Net ” consumption indicates the coal actually used per horse-power. The coal 
used (brand Kohlscheid III) contained 85.18% C, 3.49% H, 5.36% ash, 1.97% H 2 0, and 
had a mean heating value of 14 000 B.T.U. per lb. 



































































































































358 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


Table 58 

RESULTS OF TESTS WITH SUCTION GAS 


Approximate Load. 

R.p.m. 

Mean 
Effective 
Pressure, 
lbs. per 
sq.in. 

Mean 

I.H.P. 

Heating 
Value of 
Anthra¬ 
cite, 
B.T.U. 
per lb. 

Anthracite used per I.H P. Hour, 

Heat Consumed in 
Producer per 

I H P. Hour. 

tf u =14 000 B.T.U 
per lb 

H u = 14400 B.T.U. 
per lb. 

Total, 

lbs. 

Net 

lbs. 

Total, 

lbs. 

Net, 

lbs. 

Total, 

B T U. 

Net, 

BT.U. 

Nearly maximum. 

210 

108 

34.4 

14000 

.747 

.637 

.727 

.617 

10450 

8830 



(5) While Tables 57 and 58 show the results referred to the indicated horse¬ 
power, and therefore give some idea of the progress made by Giildner engines with 
regard to thermal efficiency, the tables following show the results of acceptance tests 
made by independent observers and consequently the information refers more particularly 
to the economic side. 

Table 59 

ANTHRACITE CONSUMPTION OF TWO 30-H.P. GULDNER ENGINES UNDER PARTIAL 

LOADS 

(Determined in the Municipal Lighting Plants of Mosbach and Niederbronn) 


Test 

No. 

Duration of Test 
(No. 

Interruption). 

Ap¬ 

prox¬ 

imate 

Load. 

Mean 

R.p.m. 

Effici¬ 
ency of 
Dy¬ 
namo 
and 
Belt, 

% 

Average 
Load on 
Engine. 

Me¬ 

chani¬ 

cal 

Effici¬ 

ency 

Vm> 

% 

Fuel used per Hour, 
including Coal for 
Starting, lbs. per 

Fuel used per Hour, 
excluding Coal for 
Starting, lbs. per 

Hours. 

Min. 

B.H.P. 

I.H.P. 

K.W. 

• 

B.H.P. 

I.H.P. 

K.W. 

B.H.P. 

I.H.P. 

I 

9 

20 

normal 

220.8 

85.3 

35.1 

44.2 

.80 

1.425 

.885 

.703 

1.22 

.746 

.602 

II 

8 

25 

a 

4 

220.6 

85.0 

25.9 

37.4 

.70 

1.74 

1.0 

.754 

1.43 

.897 

.618 

III 

8 

20 

* 

221 4 

82.0 

21.5 

31.1 

.69 

1.77 

1.10 

.842 

1 .56 

.945 

.740 

IV 

8 

30 

$ 

221 .9 

78.9 

17.65 

26.4 

.67 

2.93 

1.42 

.945 

1 .97 

1.13 

.753 

V 

6 

50 

i 

223.0 

72.0 

13.30 

22.2 

.60 

3.51 

1.86 

1.12 

2.83 

1.48 

.894 

VI 

7 

35 

i 

222.6 

70.2 

11.15 

18.85 

.59 

4.34 

2.23 

1.33 

3.36 

1.74 

1.03 


Maximum horse-power developed continuously by each engine, 45 for short 

periods 50-51 B.H.P. Kind of anthracite used in Tests I, II, IV, and VI (Nieder¬ 
bronn): No. Ill Nut, Portingsiepen, heating value average 13 500 B.T.U. per lb.; kind 
used in Tests III and A (Mosbach), No. Ilf Nut, Herstal (Belgium), average heating 
value 13 860 B.T.U. per lb. Efficiency of belt transmission assumed =97% in all cases, 
efficiency of dynamo determined from calibration curves of manufacturer. The fuel 
consumption is not recomputed to the basis of 14 400 B.T.U. coal. 


Table 60 


VARIATION IN SPEED IN A 30-H.P. GULDNER ENGINE IN THE MUNICIPAL LIGHTING PLANT 
AT NIEDERBRONN UNDER SUDDEN CHANGES OF LOAD 


Test 

No. 

Sudden Change of Load. 

Speed Varied. 

from 

K.W. 

to 

K.W. 

% of 
Rated 
Load. 

trom 

R.p.m. 

to 

R.p.m 

% Ot 

Normal 

Speed. 

I 

22.5 

15.6 

-32.1 

220 

223 

+ 1.50 

II 

13.4 

22.5 

+ 42.4 

225 

222 

-1.44 

III 

17.8 

8.8 

-42 

221 

225 

+ 1.68 

IV 

ib.O 

0 

-84 

221 

229 

+ 3.50 

V 

0 

20.4 

+ 95 

226 

221 

-2.48 


Remarks. 


The ^peed variation^ were observed by means 
of tachometer. 

H eight of ily-wheel appr. 5500 lbs., diameter 
9.25 ft. Coefficient of regulation at normal 
load d = 
































































STATIONARY ENGINES 


359 


Table 59 shows a comparatively slow increase of the fuel consumption per H.P. 
with decreasing load and also proves that uninterrupted continuous operation is possible 
even for the smallest loads. The latter fact, as well known, depends mainly upon the 
effectiveness of the producer end of the installation. 

The rapid accomodation of the gas producers to any calls made upon them in 
combination with the excellent speed regulation of the engine are also well brought out 
in Table 60, the figures for which are also taken from the above mentioned acceptance test. 

(c) The above mentioned Niederbronn plant tested by two Dutch engineers, Van Burkom 
and Huibers, after it had been in operation one year, at the direction of the management of 
the electric station at Scheveningen, gave the following results in everyday operation: 


Producer lighted at. 

Engine started (first ignitions). 

Time required to start installation. 

Full load on engine at. 

End of test, no interruptions. 

Time under full load. 

Total coal consumed, including that used for starting up, 

lost in ash, etc. 

Power developed, from wattmeter. 

Anthracite consumed, including coal for starting, etc., per 
effective K.W. hour. 


7:15 a.m. 

7:25 
10 min. 

7:29 
5: p.m. 

9 hrs, 31 min. 

244 lbs. 

177.2 K.W. hours. 
1.375 lbs. 


In order to be able to determine the net coal consumption, the fuel used between 
10:30 a.m. and 5 p.m., together with the electric output, were separately recorded. 
These figures gave the following results: 


Net coal used 10:30 to 5:00 . 138.5 lbs. 

Electrical output for same period. 111.3 K.W. hours. 

Hence net coal consumed per effective K.W. hour. 1.255 lbs. 

net coal consumed per I.H.P. hour, approximate.539 lbs. 


These results include all losses due to belt transmission and dynamo up to the switch¬ 
board. Coal was anthracite, Langenbrahm mine, Size II, which was used as received. 

The load on the engine during the test was kept as constant as variations in the 
power demand in the outside lines would allow. Toward the end the load was gradually 
reduced in order to let the producer burn down as far as possible. At times the mean 
effective pressure as per indicator card exceeded 100 lbs. per square inch. 

During the entire 9? hours, the producer was cleaned but once (at 2:30 in the afternoon), 
but the load was not changed, so that no interruption in the regular operation occurred. 

To complete the capacity and economy tests, a few measurements were also taken 
on the speed variations under sudden changes of load. This showed, for instance, that 
with a sudden decrease from about 35 B.H.P. to no load, the speed rose from 212 to 
220 r.p.m, which is less than 4%. 

After the completion of the tests the mixing and inlet valve was taken out for 
examination. The surfaces showed but very light deposits, not enough to interfere in 
any way with operation, although at that time the valve had been in use over fourteen 
days, from 12-14 hours each day. An examination of the cylinder walls and piston 
face also disclosed but slight deposits. According to the records of the station, the 
piston had at no time been taken out of the cylinder for purposes of cleaning, etc., 

since the engine was erected. . , ,, . 

(d) The electric station in Bergzabern, Pfalz, until 1904 operated by steam, in that 














360 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


year installed a 30-H.P. Giildner suction gas plant. The comparative statement of the 
fuel cost for the two systems is of interest. 


Table 61 


FUEL COST OF STEAM PLANT AND SUCTION GAS PLANT IN MUNICIPAL STATION AT 

BERGZABERN 


No. 

Fuel Consumption and Fuel Costs. 

Suction Gas Plant. 

Steam Plant. 

Total. 

Net,. 

Total. 

Net. 

1 

Coal consumption per B.H.P. hour, lbs. 

.848 

.720 

6.92 

6.47 

2 

Coal consumption per Kilowatt hour, lbs. 

1.475 

1.260 

12.60 

11.70 

3 

Cost of coal per B.H.P. hour, cents. 

.315 

.266 

1.32 

1.22 

4 

Cost of coal per Kilowatt hour, cents. 

.512 

.467 

2.39 

2.20 


The figures for the gas plant w r ere determined by direct measurement, the figures 
for the steam plant were computed from the average as show r n by older operating 
records of the plant. The difference between total and net consumption, that is, 
between the consumption with and without starting losses, amounted to approximately 
15% in the gas plant and 9.3% in the steam engine. At the place of consumption the 
anthracite used cost $7.35 per ton, and steam coal $3.72 per ton. 

The ratio between the fuel costs in this case is approximately 4.5 to 1 in favor of 
the gas plant. In the electric stations at Mosbach and Niederbronn, which formerly 
also operated with steam engines, the ratio is about the same (3 or 4 to 1). It must 
of course be admitted that the suction gas plant cannot show any such saving when it 
comes to compete with up-to-date and fully loaded steam plants, but even in such 
cases the economic superiority of gas over steam is certain. As far as the older 
steam plants are concerned, there is little doubt that the substitution of suction gas 
installations for steam is about the only successful means of making the operation pay. 
The above examples from practice certainly show that. 

The economy figures just quoted are all the more remarkable owing to the fact 
that they were obtained with comparatively low compression pressures. While most 
suction-gas engines operate with compression from 180 to 225 lbs., 1 some going even 
higher than that, the Giildner engine at normal load compresses to 135 lbs., and only 
compresses to from 165 to 180 lbs. w T hen working at maximum load w T ith a charge 
at larger than normal. The indicator diagrams, Figs. 

524 and 525, bring out this point, and also show 
the unusually large overload capacity of the engine. 
1 he mean effective pressure for maximum load ex¬ 
ceeds 130 lbs. per square inch. In order to get a 


P{= 16.5 lbs. per sq.in. 




Fig. 524.—Diagrams obtained at Max. Load 
and near Normal Load. 


Fig. 525. —No Load Diagram, smallest Degree of Com¬ 
pression (driving only belt and counter-shaft.) 


basis of comparison, it may be stated that in the ordinary suction-gas engine, the mean 
effective pressure is usually from 75 to 90 lbs. per square inch (see Fig. 485), and that 
only the Diesel oil engine shows anywhere near the maximum mean effective pressure 
indicated in Fig. 524. (In modern steam engines, at normal load, the mean effective 


1 Translator’s Note. These figures seem high for American practice. 

























STATIONARY ENGINES 


361 


pressure varies from 30 to 40 lbs., which is approximately \ of that shown by internal- 
combustion engines. On this basis the Giildner engine, in spite of the 4-cycle opera¬ 
tion, is about on a par with the ordinary double-acting steam engine as far as specific 
capacity is concerned.) 

7. Motorenfabrik Oberursel A. G., in Oberursel. 



Inlet valve a is automatic, ahead of this valve is placed atomizer b with overflow b'. 
At the same time that the main quantity of air is drawn through the pre-heater c, an 
auxiliary quantity is furnished through the atomizing valve. Secondary shaft d is used 
to operate alcohol pump e, magneto /, and make-and-break mechanism g„ for the 
igniter h. Exhaust valve i is controlled by the combined gear a' and a". Shaft 
governor k, by means of the rod l and the latch m, holds open the exhaust valve 
when the speed of the engine exceeds certain limits. The shift link n is used to 
reduce compression at starting. 


Table 62 

DIMENSIONS AND WEIGHTS OF “GNOM” ENGINES 


N n in B.H.P. 

1 

2 

3 

4 

5 

6 

8 

10 

12 

15 

20 

25 

R.p.m. 

400 

360 

350 

300 

300 

290 

280 

270 

200 

250 

250 

180 

Length.ft. 

Width*.ft. 

Height.ft. 

Net weight . . .lbs. 
Total weight.. lbs. 

2.30 

3.28 

3.12 

925 

1320 

2.62 

3.77 

3.45 

1515 

1925 

2.79 

3.94 

3.62 

1700 

2200 

3.11 

4.43 

3.94 

2240 

2840 

3.45 

4.60 

4.27 

2310 

2990 

3.45 

5.43 

4.43 
3380 
4180 

3.77 

5.91 

5.09 

4400 

5280 

3.94 

6.57 

5.58 
4980 
5820 

4 27 
6.38 

6 .08 
6150 
7250 

4.27 

7.38 

6.08 

7250 

8460 

4.92 

7.23 

7.23 

8340 

9450 

5.58 

6.57 

6.72 

10550 

11450 


* Parallel to shaft. 


















































































































































362 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



The engine is sold under the trade name of “ Gnom.” It may be arranged to 
operate also on kerosene or on illuminating gas by substituting proper mixing arrange¬ 
ments for the atomizer. The original designer of this machine is Hugo Seek, Sr., who 
apparently has borrowed the idea of the enclosed crank case and splash lubrication 
from American steam-engine practice. 

8. Vereit. Maschinenfabrik Augsburg und Maschinenbau-Ges. Niirnberg A.-G., 
Werk Augsburg. (See Plates XXVI to XXVIII.) 

This firm is the leading and by far the most important manufacturer of Diesel 
engines. In place of the first model, which had a cross-head, the type at present 
built operates without a cross-head (Fig. 528 and Plate XXVI) and being cheaper, finds 

a readier sale. Other licensed firms, 
however, notably the Maschinen- 
fabrik L. A. Riedinger in Augsburg, 
retain the cross-head construction, 
which undoubtedly gives a more 
satisfactory machine. 

In Plate XXVI, a is the inlet, 
b the exhaust, c the fuel injection, 
and d the starting valve, all operated 
by cams fixed on the lay shaft. 
The levers of the last two valves 
are fastened to an eccentric sleeve. 
By turning the latter through 90°, 
the two levers mentioned are shifted 
from the starting position (in which 
the fuel injection valve is out of 
commission) to the normal operating 
position (see left figure in Plate 
XXII). The starting of the engine is 
now done by operating on the 4- instead 
of the 2-cycle principle. The high 
pressure air is obtained in two stages, 
pump e through valve d taking a small 
quantity of air, preliminarily com- 
FiG. 528.— Type of small Diesel Engine. pressed, from the main cylinder near 

(Rated B.H.P.=8 at 270 r.p.m.) the end of the compression stroke, 

compressing it from about 150 to 650 
or 700 lbs. per square inch, and storing it in the starting or injection tanks. Oil 
pump /, whose suction valve is controlled by the governor, forces a quantity of oil in 
direct proportion to the load into the nozzle of the injection valve c. Piston and 
wrist-pin lubrication is taken care of by the small pump g with its oil reservoir g'. 

The operation of the engine is readily understood. During the first down stroke 
the cylinder is filled with air through the inlet valve a. This air is next compressed 
to about 480 lbs., during which it is heated to from 1000 to 1100° F. Shortly before 
the piston reaches the upper dead center, the fuel-injection valve c opens, and the oil 
which has been furnished by the oil pump /, is injected in a fine spray into the highly 
heated charge of air in the cylinder by means of high pressure air obtained as above 
described. The oil burns as it enters the cylinder, being ignited by the highly heated 
air which it meets. The expansion caused by the heat thus developed forces the power 




STATIONARY ENGINES 


363 


piston downward on the third stroke of the cycle, the pressure remaining approxi¬ 
mately constant until after about i stroke at normal load the fuel valve c closes. 
The gases then expand behind the piston to the end of the stroke. During the next 
and last stroke of the cycle the burned gases are expelled through valve b and the 
engine is then ready for the next charge. It therefore operates on the 4-cycle principle. 

In the smaller commercial machines, Fig. 528, the small air pump is attached to 
the side of the base plate at the height of the crank-shaft, and is directly driven from 
the end of this shaft. The small oil pump and the governor belonging to these 
machines are shown in Fig. 405, p. 267. 


Table 63 

DIMENSIONS AND WEIGHTS OF DIESEL ENGINES MADE BY THE MASCHINENFABRIK 

AUGSBURG 


Single-cylinder Engines. 


N n B.H.P. 

4 

6 

s 

10 

12 

15 

20 

25 

30 

35 



60 

70 








290 

280 

2 

T 0 

255 

250 

235 

215 

205 

195 

190 

180 



160 







Dimensions over all: 



















Across shaft. 

ft. 

4.27 


4.93 

5 

.25 

5.90 

6.56 

7.22 

7.55 

8.20 

8.53 

8.85 

9.50 

9.85 

10.15 

10.50 

Parallel to shaft. 

ft. 

3.28 


3.94 

3 

.94 

4.59 

4.59 

6.56 

7.05 

7.37 

7.87 

8.53 

9.18 

9.85 

10.50 

11.15 

Height above floor. 

ft. 

4.58 


5.57 

5 

.57 

6.56 

6.56 

1 7.22 

7.87 

8.53 

8.85 

9.50 

10.15 

10.80 

11.50 

12.45 

Head room required, with ref- 


















erence to dismounting 

ft. 

8.53 


9.17 

9 

.17 

9.50 

9.50 

12.15 

13.10 

13.80 

14.10 

14.45 

15.10 

16.10 

17.05 

18.05 

Depth of foundation. 

ft. 

3.28 


3.28 

3 

.94 

4.58 

5.23 

5.90 

6.23 

6.56 

6.88 

7.21 

7.53 

7.86 

8.20 

8.53 

Diameter of fly-wheel .... 

ft. 

4.27 


4.92 

5 

.23 

5.90 

6.56 

7.21 

7.53 

8.20 

8.52 

8.85 

9.50 

9.85 

10.15 

10.50 

Net weight. 

lbs. 

1980 


2640 

3190 

3850 

4520 

7140 

9680 

11900 

14300 

16900 

19600 

24200 

29700 

35200 

Total weight. 

lbs. 

2420 


3190 

3740 

4620 

5280 

8580 

11450 

14300 

17200 

20200 

23300 

28600 

35200 

41800 



Single-cylin¬ 
der Engines. 

Twin-cylinder Engines. 

B.H.P. 


80 


100 

30 


40 

50 

60 

70 

80 

100 

120 

140 

160 

200 













R p.m. 

160 

160 

235 


215 

205 

195 

190 

180 

170 

165 

160 

160 

160 















Dimensions over all: 



















Across shaft. 

. ft. 

11.15 

11. 

50 

7 

.21 

7.54 

8.21 

8.53 

8.85 

9.50 

9.85 

10.15 

10.50 

11.15 

11.50 

Parallel to shaft. 

. ft. 

11.80 

12. 

45 

9 

.20 

9.85 

10.80 

11.15 

11.80 

12.45 

13.10 

14.10 

15.10 

15.70 

16.40 

Height above floor. 

. ft. 

13.45 

14. 

45 

7 

.20 

7.86 

8.53 

8.85 

9.50 

10.15 

10.80 

11.50 

12.45 

13.45 

14.45 

Head room required, with ref- 


















erence to dismounting 

. ft. 

19.70 

21. 

3 5 

12 

.45 

13.10 

13.80 

14.10 

14.45 

15.10 

16.05 

17.05 

18.05 

19.70 

21.35 

Depth of foundation. 

. ft. 

8.85 

9. 

20 

5 

.91 

6.23 

6.56 

6.87 

7.21 

7.53 

7.86 

8.20 

8.52 

8.85 

9.1 7 

Diameter of fly-wheel . . . . 

. ft. 

11.15 

11. 

50 

7 

.21 


7. 53 

8.20 

8.53 

8.85 

9.50 

9.85 

10.15 

10.80 

11.15 

11.50 

Net weight. 

lbs. 

40700 

50600 

12800 

16900 

21100 

25500 

30100 

35200 

44000 

54800 

63800 

72600 

91300 

Total weight. 

lbs. 

47200 

59400 

15400 

20200 

25300 

30800 

35800 

41800 

51600 

63800 

74800 

84600 

105500 


Operating Results with the Diesel engines made by the Maschinenfabrik Augsburg, 
(a) Tests made by Prof. E. Meyer, in September, 1900, on a 30-H.P. motor of the 
older type (with cross-head) but compressing the injection air in two stages, as in later 
types: Principal dimensions of the working cylinder, .0 = 11.8", S = 18.25"; of the air 
pump, 0 = 1.97", 5 = 3.15"; ratio of compression, £ = 16.3. The first six tests recorded 
in Table 64 were made with American kerosene, ^ = .796 at 64.4° F., the last four with 
Tegernseer crude oil, 7 = .789 at 68.4° F. 




































































364 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 

Table 64 



177.4 

181.1 

182.6 

184.0 

183.3 

185.8 

181.6 

181.2 

181.8 

185.0 

Brake horse-power. 

Mean effective pressure in power.. 

38.90 

29.8 

30.0 

23.45 

15.05 

8.41 

29.8 

29.75 

23.18 

15.18 

cylinder, lbs. per sq.in. 

Mean indicated friction resistance, 

106.0 

85.4 

85.4 

70.3 

53.2 

37.9 

85.7 

88.2 

71.0 

55.9 

p r , lbs. per sq.in. 

Indicated horse-power *. 

19.25 

47.50 

20.0 

38.90 

20.0 

39.20 

19.85 

32.60 

20.7 

24.70 

20.0 

17.75 

39.20 

40.30 

32.55 

26.00 

Mechanical efficiency, . % 

81.9 

76.3 

76.3 

71.9 

61.0 

47.3 

76.0 

73.7 

71.2 

58.3 

Oil used per I.H.P. hour .. . .lbs. 

.402 

.348 

.346 

.348 

.350 

.366 

.357 

.350 

.352 

.335 

Oil used per B.H.P. hour. .. .Ibo. 

.482 

.457 

.454 

.483 

.576 

.764 

.469 

.475 

.496 

.575 


* Without deduction for air pump horse-power, but less an allowance equivalent to 1.4 lbs. per sq.in. for 
suction and exhaust resistances. 


Assuming that the heating value of the oil is H u = 18 550 B.T.U. per lb., the above 
figures show the indicated thermal efficiency at full load to be rji = approximately 39 % 
and the economic efficiency r) w = approximately 30%. 

( 6 ) Following tests (a), Prof. Meyer also determined the fuel consumption of a 
Diesel engine of the older type, which had been in operation in a textile mill since the 
beginning of 1899. The engine was of the two-cylinder type, 60 B.H.P., with cross-head 
and single-stage air pump. It received no special preparation for the test. The oil 
consumption amounted to .407 lb. per I.H.P. hour at 68.3 I.H.P. (after subtracting 
pump work). The oil used was Tegernseer crude ( 7 * = .8125 at 62.6° F.). Estimating 
the mechanical efficiency at i) m = 69.5%, this means a consumption of .513 lb. of oil 
per B.H.P. hour and an economic efficiency of approximately 26%. 

(c) The thermal efficiency of the newer design of Diesel engine, without cross-head, 
is shown by a series of tests made by Prof. E. Meyer in June, 1902, on two engines of 
/0 and 8 B.H.P. respectively. The details of the 70 B.H.P. engine are shown in Plate 
XX'V I, those of the 8 B.H.P. machine in Fig. 528. Table 65 contains the important 
results of the tests. The engine dimensions y^ere: 


Nn = 70 B.H.P. { 


Power cyl. D = 15.75", 
Pump cyl. D= 2.21", 


Nn 


B.H.P. 


Power cyl. D = 6.48", 
Pump cyl. D= .59", 


*8 = 23.60", 
* 8 = 5.47", 

*8=10.70", 
*S= 1.97", 


£ = 15.4 

Air pressure at beg. = 120-150 lbs. 
£ = ? 

Air pressure at beg. = 120-150 lbs. 


The fuel used for the first five tests on the large engine was Russian kerosene 
(r = .806 at 65.5° ), for the rest of the trials the fuel was a paraffin oil of dirty-brown 
color ( r = .893 at 59°). The small engine was tested only with the Russian kerosene. 

The heating value of the kerosene was afterward determined with great accuracy at 
18 550 B.T.U. per pound, consequently for the tests showing the best oil consumption 
in the table the thermal efficiencies figure out as follows: 


70 B.H.P. Engine. 8 B.H.P. Engine. 

Indicated thermal efficiency, jji.402 353 

Economic efficiency, t) w . 320 280 

Both of these engines obtained their air for injection purposes by preliminary 
compression in the main cylinder, as above explained. The scheme has certain disad¬ 
vantages, for which reason several of the licensed firms have already abandoned its use 



























STATIONARY ENGINES 


365 


Table 65 


70 B.H.P. Engine. 


Revolutions per minute. 

157.9 

158.8 

158.7 

159.8 

160.5 

157.8 

159.0 

159.3 

159.9 

B.H.P. 

Mean effective pressure power 

85.3 

68.6 

68.55 

52.2 

34.40 

85.25 

68.70 

68.85 

52.25 

cylinder, lbs. per sq.in. 

116.50 

96.20 

96.50 

79.00 

57.60 

122.00 

101.00 

101.80 

82.80 

Total I.H.P. 

Mean effective pressure, air 

107.5 

89.2 

89.3 

73.5 

54.0 

112.3 

94.1 

95.0 

77.5 

pump, lbs. per sq.in. 

Indicated pump H.P. (esti- 

304.00 

277.00 

274.00 

295.50 

226.75 

288.00 

309.50 

315.00 

316.50 

mated for the small engine).. 

2.53 

2.32 

2.29 

2.48 

1.93 

2.40 

2.60 

2.67 

2.68 

Net I.H.P. 

105.2 

86.9 

87.1 

71.0 

52.0 

109.9 

91.5 

92.3 

74.9 

Mechanical efficiency.% 

81.4 

79.1 

78.9 

73.6 

66.2 

77.7 

75.2 

74.8 

70.0 

Oil used per I.H.P. hour . . . lbs. 

.342 

.339 

.339 

.328 

.331 

.361 

.339 

.344 

.335 

Oil used per B.H.P. hour. . . lbs. 
Cooling water per hour (inlet 

.420 

.430 

.432 

.450 

.500 

.467 

.455 

.460 

.479 

temp. = 51.8° F.).lbs. 

Temperature of cooling water, 

— 

1275 

1945 

925 

308 

2240 

1540 

1495 

1715 

outlet. 0 F. 

Exhaust temp, (measured in 

152 

162-178 

155-158 

158-170 

160-173 

171-158 

165-168 

155-175 

171-175 

pipe just beyond valve) ° F. .. 

874 

637-660 

644-660 

635-637 

462-457 

787-792 

652-682 

627-719 610-627 



8 B.H.P. Engine. 

Revolutions per minute. 

267.1 

278.4 

270.3 

274.7 

276.3 

284 

B.H.P. 

9.91 

8.75 

S .49 

6.13 

4.61 

0 

Mean effective pressure power cylinder, lbs. per sq.in. 

107.50 

92.80 

94.30 

75.15 

61.60 

26.95 

Total I.H.P. 

12.76 

11.50 

11.32 

9.20 

7.55 

3.40 

Mean effective pressure, air pump, lbs. per sq.in. 

— 

— 

— 

— 

— 

— 

Indicated pump H.P. (estimated). 

.30 

.30 

.30 

.30 

.25 

.20 

Net I.H.P. 

12.46 

11.20 

11.02 

8.90 

7.30 

3.20 

Mechanical efficiency.% 

79.4 

78.1 

77.0 

69.0 

63.2 

— 

Oil used per I.H.P. hour.lbs. 

.389 

.397 

.384 

.344 

.366 

— 

Oil used per B.H.P. hour.lbs. 

.489 

.507 

.495 

.522 

.581 

— 

Cooling water per hour (inlet temp. = 51.8° F.).lbs. 

216 

297 

— 

— 

— 

— 

Temperature of cooling water, outlet.° F. 

167 

145 

— 

— 

— 

— 

Exhaust temp, (measured in pipe just beyond valve ).°F. 








( d ) Worthy of attention, also, are a series of tachometer measurements which the 
Maschinenfabrik Augsburg itself has made on an 80-H.P. two-cylinder engine, having 
cross-head and single-stage air pump. The dimensions of the engine were: Z) = 11.80", 
stroke = 18.10", n = 190 r.p.m. It was direct connected to a dynamo and had two 
fly-wheels, the weight of each of which was 5720 lbs., with an effective radius of 
3.68 ft. The computed coefficient of regulation at full load was 1% (d = .01). The 
governor tvas adjusted to regulate to within 4%. The measurements were taken with a 
Horn tachograph, the diagram showing the following results: 

At the normal load the tachograph indicated a value of d = when one wheel was 
in use, and a value of d = when both wheels were in place. Both values, however, 
are larger than the value of d computed on the basis of fly-wheel weight. This is 
probably due to inaccuracies in the tachograph diagrams, owing to inertia effects in the 
moving parts of the instrument itself. The voltmeter for all changes of load, whether 
one or both wheels were used, did not show a variation of more than \ volt. A 





















































366 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


25 candle-power incandescent lamp showed a slightly unsteady light when the engine 
was running at about normal load with one wheel; the light, how r ever, was absolutely 
steady when two wdieels were employed. 


Table 66 

SPEED VARIATION IN AN 80-H.P. DIESEL ENGINE 


Only One Wheel in Use. 

Both W 7 heels in Use. 

Sudden Change of Load 

Per Cent of 

Speed 
Variation, 
Per Cent. 

Sudden Change of Load 

Per Cent of 

Speed 

Variation, 

Per Cent. 

from K.W. 

to K.W 7 . 

Normal Load. 

from K.W 7 . 

to K.W. 

Normal Load. 

24.2 

42.9 

+ 48.0 

- 2.2 

41.7 

23.8 

-46.0 

+ 1.73 

43.2 

24.2 

-48.8 

+ 2.25 

24.0 

16.75 

-18.6 

+ .51 

24.2 

16.9 

-18.7 

+ 1.0 

16.75 

8.35 

-21.5 

+ .60 

16.9 

8.14 

-22.5 

+ 1.15 

16.5 

23.7 

+ 18.5 

- .75 

8.14 

24.2 

+ 41.2 

- 2.0 

23.7 

42.2 

+ 47.5 

- 2.0 

24.2 

43.2 

+ 48.8 

-2.4 

42.2 

16.5 

- 66.1 

+ 2.4 

43.2 

16.9 

-67.5 

+ 2.95 

16.5 

41.8 

+ 64.9 

-2.1 





41.8 

7.9 

-87.0 

+ 3.75 


Table 67 

FUEL CONSUMPTION OF DIESEL ENGINES PER B.H.P. HOUR, HEATING VALUE OF THE 

OIL ASSUMED 18 000 B.T.U. PER POUND 

(Guarantee figures furnished by the Maschinenfabrik Augsburg, reserving a permissible fluctuation of 10 %.) 


Sinsrle-cvlinder ene-ine. 

HP 

8 

10 

12 

15 

20 

25 

30 

35 


Two-cylinder engine. 

.H.P. 

.. ! .. 


30 

40 

50 

60 

70 

From normal to full load .. . 

/lbs. 

.517 

.507 

.485 

.473 

.462 

.452 

.440 

.430 

' * * \ B.T.U. 

9300 

9120 

8730 

8520 

8300 

8120 

7920 

7750 

Three-quarter load .. 

/ lbs. 

• • ' \ B.T.U. 

.538 

9700 

.528 

9500 

.507 

9120 

.495 

8920 

.485 

8730 

.473 

8520 

.462 

8300 

.452 

8120 

One-half load. 

[lbs. 

.627 

.616 

.594 

.573 

.560 

.550 

.538 

.528 


11250 

11080 

10690 

10300 

10080 

9900 

9700 

9500 

One-quarter load. 

/ lbs. 

.858 

.825 

.792 

.770 

.737 

.727 

.715 

.705 

' • • \ B.T.U. 

15400 

14830 

14250 

13850 

13250 

13050 

12880 

12690 


Single-cylinder engine. 


40 

50 

60 

70 

80 

100 

125 

Two-cylinder engine. 


80 

100 

120 

140 

160 

200 

250 

From normal to full load. 

/ lbs. 

.430 

.430 

.418 

.407 

.407 

.407 

.407 


.1 B.T.U. 

7750 

7750 

7530 

7320 

7320 

7320 

7320 

Three-quarter load. 

/ lbs. 

.452 

.452 

.440 

.430 

.430 

.430 

.430 


.\ B.T.U. 

8120 

8120 

7920 

7750 

7750 

7750 

7750 

One-half load.. 

/lbs. 

.517 

.517 

.507 

.495 

.495 

.495 

.495 



9300 

9300 

9120 

8920 

8920 

8920 

8920 

One-quarter load. 

/ lbs. 

.694 

.683 

.672 

.660 

.660 

.660 

.660 


.\ B.T.U. 

12500 

12300 

12080 

11880 

11880 

11880 

11880 








































































PLATE XXVII 



Key: 

e= Lubricating oil filter; 
/= Jacket water pumps; 
g = Exhaust muffler. 


Key: 

a = Fuel oil filter; 

fr = Fuel injection air tanks; 

c— Starting air tanks; 

d = Collecting tank for waste lubricating oil; 


Plan of a Diesel Oil Engine Installation, Total Capacity 400 H.P. 

(3 Engines of 70 H.P. each, 1 Engine of 200 H.P.) 

Built by the Vereinigte Maschinenfabrik Augsburg und Machinenbaugesellschaft Niirnberg A.-G., Werk 

Augsburg. 














































































































































































































































































































































































































































































































































































































































































































































































































































Plan of a Diesel Oil Engine Installation. Total Capacity 

1600 H.P. 

(Four Engines, each 400 B.H.P.) 

Built by the Vereinigte Maschinenfabrik Augsburg und Maschinenbau- 
gesellschaft Nurnberg A.-G., Werk Augsburg. 


Key: 

a = Fuel storage tank (each 105 cu.ft. capacity); 

5 = Cooling water storage tanks (each 350 cu.ft. capacity); 
c= Lubricating oil tanks (each 150 gallons capacity) ; 
d = Fuel oil alter; 
e = Starting air tanks; 

/= Fuel injection air tanks; 
g = Exhaust mufflers; 

h = Return pipes and collecting tanks for waste lubricating oil. 


A 























































































































































































































































































































































































































































































































































































































































































STATIONARY ENGINES 


367 


III. Other Types of Foreign Gas Engines 

1. Langen & Wolf in Wien (Vienna). This firm is the Austrian licensee of the 
Gasmotoren-Fabrik Deutz, but has in the past few years brought out a construction 
of its own, the design of H. Ebbs, one of the partners of the firm. An external view 
of this machine is shown in Fig. 529. 

The engine has a pleasing appearance. The construction is especially distinguished 
by the two distance pieces, before the first cylinder and between the cylinders, and the 



Fig. 529— General View of a 300 H P Double-acting Tandem Engine,Langen & Wolf, Vienna. 


method of tying together frame, cylinders, and distance pieces by four continuous 
tie-rods or bolts. The latter are intended to insure central alignment of the 
various parts and to relieve the cast-iron distance pieces of tensile stresses as far as 
possible. 

Further details of construction are given by the cross-section, Fig. 530, p. 368. 
The back cylinder cover is easily dismounted and is made of such diameter that the 
piston may be taken out at that end without trouble. The jacket wall is cast in one 
piece with the cylinder, and in this connection the division of the wall at the middle 
plane a for the purpose of relieving temperature stresses is worthy of attention. The 
engine is intended to operate on producer gas, but to make possible the use of 
illuminating gas, where the latter may be used as a reserve or as the starting 
medium, the inlet valve housings are supplied with suitable connecting flanges for 
this purpose. For information regarding valve gear and regulations, see Fig. 277, 

2. Societe anon. John Cockerill in Seraing. (See also Plates XXIX and XXX.) 
The engines built by this company all operate on the ordinary 4-cycle principle, 
and therefore need no very extended description. The various types at present 
brought out are well illustrated in the following figures and plates. (Of special 
interest is the new valve gear employing eccentrics, for details of which see bigs. 
278-282.) 


i 






368 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 






Fig. 530.—Cross-section of a 100-120 H.P. Double-acting 4-cycle Engine. Langen & Wolf, Vienna. (For Valve Gear, see Fig. 277.) 



















































































































STATIONARY ENGINES 


369 


Operating Results . 1 (a) Tests made Witz and Hubert, 1898, on one of the older 
200-H.P. simplex engines built by the Cockerill Co. for blast-furnace gas. The cylinder 
diameter was 31.5", the stroke 39.3". The test lasted twenty-four hours: 


R -P-m. 105.2 

Explosions per minute. 47.0 

. 179.00 

I.H.P., approximate. 212.00 

Mechanical efficiency. 85% 

Gas used per B.H.P. hour. 119.1 eu.ft. 

Higher heating value of gas. 110 B.T.U. 

Water used in the washers, per eu.ft of gas. 19.8 lbs. 

per B.H.P. hour. 67.0 lbs. 

Cooling water per B.H.P. hour . . . .:. 93.3 lbs. 

Temperature cooling water, inlet. 62.8° F. 

outlet. 93.6° F. 

Total water consumption per B.H.P. hour. 160.3 lbs. 

Cylinder oil per B.H.P. hour.033 lb. 

Grease for bearings per B.H.P. hour. .0042 lb. 


The compression was 107 lbs. per sq.in., the mean effective pressure 52.5 lbs. per sq.in. 
Gas consumption and heating value are referred to 46.4° F. and 760 mm. Hg. 



Fig. 531.—End View, Cockerill Engine. (Belongs to Plate XXIX.) 


1 Translator’s Note. The tests quoted under (a) and ( b ) are given in the historical part of 
Giildner’s Handbook, not translated in this volume. They are quoted here to show the results 
attained with the early Simplex engines, designed by Delamare-Deboutteville, and built by the 
Cockerill Co. 
































































370 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


(6) Tests made by Hubert early in 1900, on a single-acting 600-H.P. simplex 
blast-furnace gas engine built by the Cockerill Co. Cylinder diameter 51.2", stroke 55": 


Test I. Engine without Blowing Cylinder (6 hours) 


R.p.m. 94.57 

Explosions per minute . . ;. 41.90 

I.H.P. 775.0 

B.H.P. 567.0 

Mechanical efficiency, %. 73.14 

Gas used per B.H.P. hour, cu.ft. 124.8 

Heating value of gas, B.T.U. per cu.ft. 110.5 

Temperature of gas at cylinder, °F. 48.2 

Temperature of exhaust gas, °F. 948.0 

Cooling water used per B.H.P. hour: In cylinder jacket, lbs. . 126.50 

In piston, lbs. 28.30 

Total cooling water, per B.H.P. hour. 154.80 

Temperature cooling water, inlet, °F. 46.15 

outlet, °F. 91.70 


Tests II and III. Blowing Cylinder in Operation. (Each test 2 hours) 


II. III. 

R.p.m. 83.92 93.02 

I.H.P., engine. 736.00 873.00 

I.H.P. in blowing cylinder. 555.0 715.00 

Total mechanical efficiency. 75.33 81.81 

Gas consumption per B.H.P. hour, cu.ft. 112.0 101.8 

Heating value of gas, av. B.T.U. per cu.ft. 111.2 112.7 

Gas temperature at cylinder, °F. 46.75 49.80 

Temperature of exhaust gas, °F. 912 1017 

Cooling water per B.H.P. hour: Cyl. jacket, lbs... 115.5 89.3 

Piston, lbs. 29.9 22.7 

Total, lbs. 145.4 112.0 

Temperature increase of cooling water, °F. 82.10 87.60 


The average heat balance was as follows: 28% transformed into work, 52% lost in 
cooling water, and 20% lost in exhaust and by radiation. The compression pressure 
was 135 lbs. per square inch, the mean effective pressure 67.4 lbs. per square inch, the 
explosion pressure from 248 to 284 lbs. The heating values of the gas above given is 
referred to 46.4° F. and 760 mm. Hg., and was determined by Witz in a bomb calorim¬ 
eter. Tests with Junker’s calorimeter gave only 88, 102.5, and 100 B.T.U. per cubic 
foot, that is, approximately 10% less. 

(c) In addition to the above the following tabic shows a series of results obtained 
by Prof. Hubert in 1901 on tests of a 200-H.P. Cockerill engine: 1 


Principal Dimensions 

Cylinder diameter. 33.45", Stroke . . 

Normal r.p.m. 

Diameter of fly-wheel, feet. 

Peripheral velocity of wheel ft. per sec. 

Weight of wheel, lbs. 

Fly-wheel weight per H.P. approximate, lbs. . . 


39.40" 

100 

13.10 

69.0 

23000 

165.0 


1 The original report published in the Revue univ. des mines, etc., 1902, p. 273, contains numerous 
diagrams and tachographs. 


































’LATE XXIX 


t 





SxAaosi /rrlet 
Va/i/e Va/xe 


For Detail of Valve 
Gear, see Figs. 
278-282. 


— 


; t— 




e 

@ rta @ 

fe 1 

- 0 ) 


■- 

— o — 




i§ 

® m © 

i> 

J 



—1___1— 









S.I I ' 

-v / 




(1^ 



— 






1 


mFT 7 




=u=ii=gin if h 



•1/ 








_ 


Assembly Drawing, 3000 B.H.P. Double=acting 4=cycle Tandem Engine. 


Built by Soc. An. John Cockerill, Seraing (Belgium). 




















































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































PLATE XXX 






■ mm/mw/M-/,, 


To con tar of next orut- 
of tiooBH/ 9 - 3* 4 ft 


/(<?y to Colors 


\AtrMaL7-L 


Exhaust dam 


nrv 


p— 





J- - - - T T - 

, f 



r 




Erecting and Piping Pian of the 3600 H.P. Central Station, Soc. John Cockerill, Seraing, Belgium. 


(Simplex Blast-furnace Gas Engine, 600 and 1200 II. 1 . Rated ( apacity.) 



























































































































































































































































































































































































































































































































































































































































































































































































































































/ffS.4 ——- f— - - - £0»0 


STATIONARY ENGINES 


371 
























































































































































































































372 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Fig. 533.—General View of a Double-acting 4-Cycle Cockerill Engine direct-connected to Blowing Cylinder. 
Rated H.P. = 15C0 at 80 r.p.m. (D= 51.2", S = 54.2".) 


















STATIONARY ENGINES 


373 



Fig. 534.—General View of a Double-acting Tandem Cockerill Engine. 
Rated H.P. = 300 at 80 r.p.m. (D = 51.2", >8 = 54.2".) 






























374 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


Table 68 


Test 

No. 

Load. 

R.p.m. 

I.H.P. 

B.H.P. 

Me¬ 

chanical 

Effici¬ 

ency, 

%. 

Heating 
V alue 
of Gas, 
B.T.U. 

per 

cu.ft. 

Gas Used 
per 

Heat Supplied 
per 

Indi¬ 

cated 

Ther¬ 

mal 

Effici¬ 

ency, 

%. 

Eco¬ 

nomic 

Effici¬ 

ency, 

%. 

Heat in 
Cooling 
W ater, 

%. 

Temper¬ 
ature of 
Exhaust 
Gas, 
Deg. F. 

I.H.P. 

Hour, 

cu.ft. 

B.H.P. 

Hour, 

cu.ft. 

I.H.P. 

Hour, 

B.T.U. 

B.H.P. 

Hour, 

B.T.U. 

1 

no load 

101.35 

45.5 

0 


108.9 

247.5 


27000 


9.4 


30.0 

915 

2 

i 

103.20 

90.0 

63.0 

70.2 

116.0 

174.5 

247.0 

20900 

28620 

12.6 

8.9 

22.1 

955 

3 

i 

98.56 

152.5 

121.9 

80.0 

118.2 

115.5 

144.5 

13640 

17050 

18.6 

14.9 

29.4 

985 

4 

i 

98.90 

208.5 

173.0 

82.9 

113.0 

97.3 

117.2 

11000 

13250 

23.2 

19.1 

26.7 

1042 

5 

full 

98.60 

224.0 

196.9 

87.6 

113.8 

115.0 

131.3 

13100 

14950 

19.5 

17.0 

30.2 

1192 

6 

full 

98.60 

242.5 

218.0 

90.0 

137.2 

108.7 

120.8 

12440 

13750 

20.5 

18.5 

30.0 

1164 

7 

no load 

105.90 

46.8 

0 

_ 

110.2 

244.5 

_ 

27000 

— 

9.4 

— 

47.4 

916 

8 

\ 

106.35 

106.2 

56.5 

53.3 

101.7 

141.2 

265.0 

14320 

26900 

17.8 

9.5 

36.1 

962 

9 

h 

100.62 

150.5 

110.8 

73.8 

96.2 

107.5 

146.0 

10230 

13950 

24.8 

18.2 

44.0 

995 

10 

l 

100.37 

206.0 

167.8 

81.5 

98.9 

107.5 

131.9 

10980 

13080 

23.2 

19.3 

31 .3 

1037 

11 

full 

98.18 

233.0 

209.0 

90.0 

109.5 

106.5 

119.0 

11670 

12980 

21.9 

19.6 

30.9 

1092 

12 

no load 

101.37 

57.0 

0 

_ 

103.8 

270.0 

_ 

28000 

— 

9.1 

— 

26.3 

785 

13 

i 

101.33 

100.3 

53.7 

53.6 

100.8 

179.5 

334.0 

18050 

33700 

14.1 

7.6 

28.3 

924 

14 

i 

100.90 

149.5 

111.0 

74.5 

101.1 

136.0 

182.0 

13900 

17630 

18.4 

14.5 

26.0 

970 

15 

2 

100.00 

202.3 

167.0 

82.8 

96.4 

114.0 

137.1 

11200 

13300 

23.1 

19.1 

27.7 

1029 

16 

full 

99.64 

238.0 

212.0 

89.2 

108.9 

110.2 

124.0 

12100 

13440 

21.3 

18.9 

23.3 

1092 

17 

full 

99.74 

244.0 

212.3 

87.2 

104.1 

106.5 

122.2 

11100 

12720| 23.0 

1 

20.0 

24.2 

1107 


For tests Nos. 1 to 11 the engine was regulated by means of a pneumatic governor; 
for the last six tests an ordinary centrifugal governor was used to control the speed. 
With either governor the speed variation under sudden increase or decrease of load 
varying between normal load and no load was from 4 to 7%, the total speed variation 
between these two limits of load being on the average 10%. Within one 4-stroke 
cycle, the angular velocity at normal loads varied from 3 to 4%, at the maximum 
load from 5 to 6%. The average compression and explosion pressures w T ere respect¬ 
ively: At ^ load, 55.5 and 105 lbs. per square inch; at \ load, 61.2 and 156 lbs. per 
square inch; at f load, 119 and 191 lbs. per square inch; and at maximum load, 153.5 
and 239 lbs. per square inch. 

3. Compagnie “Duplex” pour la fabrication des moteurs a gas in Paris. 

One of the first French engine builders to use the double-acting 4-cycle engine for large 
capacities was Niel, the designer of the “ Duplex ” engine. The machine does not offer 
much of anything new regarding construction, but employs a method of charging which 
is quite different from that used in other engines. The charge is drawn in on one side 
of the piston and is distributed between the two cylinder ends, so that each end works 
with only about one-half full charge for the purpose of obtaining complete expansion. 

As shown in Figs. 535-536, 1 each cylinder end has an inlet valve b and an exhaust 
valve a. The two inlet valves are connected by a passage or port c, which at the 
middle carries the mixing valve d, which in turn obtains air through the valve e and 
gas through valve /. All the valves are operated by cams from a lay shaft. The 
operation is as follows: In Fig. 536, as the piston starts to the right it draws in a full 
charge behind it, at the same time forcing the exhaust gas out of the right-hand end 


1 Z. d. V. D. I., 1901, p. 325. 




































STATIONARY ENGINES 


375 


of the cylinder. On the return stroke one-half of the charge in the left-hand end is forced 
through the port c and the inlet valves into the right-hand end of the cylinder, the 




valves d, e, and / having closed. At the middle of the stroke, both valves b also 
close. Then follows compression in the left end and expansion of the charge in the 
right end of the cylinder. The charge in the left end is ignited by hot tube at the 
proper time. The piston now moving to the 
right compresses the mixture in that end, 
this is in turn ignited when the piston reaches the 
crank end. The two ignitions are therefore 180° 
apart, while exhaust and charging actions occupy 
the other 540° of the cycle. When the speed ex¬ 
ceeds the normal, the governor temporarily pre¬ 
vents the opening of the gas valve /. 

The indicator cards taken when operating 
with illuminating gas show that the compression 
is about 90 lbs. and the explosion pressure varies 
from 300 to 450 lbs. The mean effective pressure 
is only about 43 to 50 lbs. per square inch. The 
terminal pressure is close to atmosphere, so 
that the exhaust is practically noiseless. Definite 

test figures for this engine are apparently not available, which does not speak particu- 
larly well for its economy. 



Fig. 537. 


























































































































376 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


4. Societe anon. (Sexploitation des brevets Letombe in Lille. This engine, which 
received a great deal of attention at the last world’s fair at Paris, is distinguished by a 
very promising system of governing, which regulates both quantity and quality of 
charge. In contradistinction to other methods of governing, the quantity of the charge 
decreases and the gas content increases as the load increases, that is, the higher the 



Figs. 538 and 539.—Letombe Engine. 


load, the smaller the total volume of charge, but the richer the mixture. On the 
other hand, in accord with reasons based on thermal considerations, the leaner mixtures 
at the lower loads receive the higher compression. 

For larger capacities, this engine is built as a double-acting 4-cycle, Figs. 538 and 
539. The most important details of its design are found in the arrangement and 
operation of the air and gas valves. The main inlet valve a, Figs. 540 and 541, is 



Figs. 540 and 541. 


automatic, that is, it acts as a simple check or suction valve. The governing valve b , 
placed just ahead of a, is actuated by the step cam c, the valve stem of b also lifting 
the gas valve d. The governor controls the position of the roller e, so that it runs on 
one of the four steps of the cam, depending upon the load. Valve b is opened at the 
same time for all loads, but the moment of closing and the effective lift of this valve 
are so regulated that, if the load is decreasing, the valve closes later and the lift is 











































































STATIONARY ENGINES 


377 


smaller than it was before. Since gas valve d is directly operated by b, this valve is 
affected in the same way. The last step of the cone is so low that the gas valve d is 
not lifted at all, in which case the engine will miss the next 
cycle. 

In the diagram, Fig. 542, line ab is the compression line for maxi¬ 
mum load iV max (charge =50-55% cylinder volume, compression =120 
lbs.), while line cd is the compression line for N min (charge = cylinder 
volume, and compression = 195 lbs.). The work area developed is there¬ 
fore controlled by shifting the compression line, while expansion line 
and explosion pressure remain approximately the same. 

Operating Results. The first accurate tests of an engine of this 
type were made by Prof. Witz in July, 1902, on a tandem engine, in 
which the front cylinder was single-acting, the back cylinder double¬ 
acting. The engine was rated at 300 B.H.P. and was intended to operate 
on producer gas. Air and steam were forced through producer and 
the gas made through scrubber to a gas holder 
by means of a Root blower, which was operated 
by a 3-H.P. auxiliary engine. The cylinder 
diameter of the main engine was 23.65", the 
stroke was 31.5". The piston rod between 
the two cylinders had a diameter of 7.08". 
The coal used was anthracite containing 14.8% 
of volatile matter and about 3% of ash. The 
test lasted eleven hours, with an interruption of one hour, and gave the following 
average results: 

Table 69 

R.p.m. 135.56 

Total B.H.P... 289.00 

H.P. in blower engine. 3.00 

Net B.H.P. 286.00 

Anthracite per B.H.P. hour, lbs. -830 

Heating value of gas per bomb calorimeter, B.T.U. per cu.ft.. 148.00 

Temperature of cooling water, °F.131-149 

Temperature of exhaust gas, °F.... 658 

Average compression pressure, lbs. per sq.in. 118 



4ig. 542. 


The economy is apparently very good, but no better than is attained by other 
first class machines regulating by means of one or the other of the common systems 
of governing. Compare, for instance, pp. 328 and 358. 

5. Premier Gas Engine Co., Nottingham. The Premier gas engine, introduced 
during the last few years for large capacities, apparently with great success, is of 
the 4-cycle positive scavenging type, that is, near the end of exhaust stroke the com¬ 
bustion space is thoroughly flushed out with fresh air in order to receive a fresh 
cool charge and to cool down the interior cylinder and piston walls. 1 Accurate 
drawings of the constructive details of the engine were not obtainable, but the 


1 This method of scavenging a 4-cycle cylinder before the new charge comes in is not a new idea 
and was already in use by Daimler in 1885 in some of his small engines, but the Premier Co. first 
applied it to large engines. 
















378 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


method of operation and the general features of the design may be made clear by 
Fig. 543, which represents a 600-H.P. tandem engine. 1 All Premier engines less than 
300 B.H.P. have only one cylinder. 

Air pump a, beginning with the crank position shown in the figure (about .8 
of the exhaust stroke), scavenges alternately the combustion spaces b and c, until 
the end of the stroke. The air supply in each case amounts to about 3£ times 

the volume of the space to be cleared out, and passes through ports d and d', and 

inlet valves e and e ' respectively. At the end of the scavenging period in the 
cylinder concerned, the suction stroke next follows as in ordinary operation. The 
power pistons b and c are connected by means of side rods /, which on one end 
are held by the yoke g, while on the other they are connected to the pump piston 

a. The rods thus pass through the annular space of the pump a, but not through 

the combustion space of the cylinder b. Grid valve h is operated by rod and eccentric 
from the main shaft, and not only serves to control the suction and discharge passages 



Fig. 543. —Premier Engine. 


of the pump a, but also at the proper time establishes communication between the 
outer air and the passage d through which one or the other of the power cylinders 
may then draw its charge of air. During the first third of the instroke, pump a 
forces a part of its charge of air back through the suction ports, for the rest of 

the stroke up to the beginning of scavenging the air is compressed into the ports d 

and d . Each inlet valve housing e is furnished with a special piston valve which 

keeps the gas ports closed during scavenging, but opens the air ports wide. During 

the suction stroke this valve then opens the gas ports and closes the air ports partly 
until the right ratio of air to gas is obtained. When the speed becomes too high 
the governor throws the operating rods of the gas valve temporarily out of action. 
The gas ports then remain closed and the cylinder draws nothing but air. The 
heads of the power piston as well as the exhaust valves are water-cooled. Ignition 
is effected by two open hot tubes in each cylinder. 

Operating Results. (a) Professor Humphrey obtained the following results on a 
6 0°-H P. tandem engine operating on Mond gas. The machine had a cylinder diameter 
of 28 , a stroke of 30", while the speed was 128 r.p.m. 


Total indicated horse-power, 489 
Indicated pump horse-power, 34 


Net I.H.P., 455 


Engineering, 1901, p. 195. 































STATIONARY ENGINES 


379 


Brake horse-power, 386.0 

Total mechanical efficiency, ij m = 75.7%, net ijm=81.0%. 

Heating value of Mond gas, approximate, 144 B.T.U. per cu.ft. ] rji =31.6% 

Gas used per B.H.P. hour, 69.2 cu.ft. I^ = 25.6% 

In this case both cylinders worked at about § load each. 

(b) A second test on the same engine, made by loading up the back cylinder and 
cutting out the front cylinder, gave results as follows: 

n = 127.4 r.p.m., 'N% = 320.92 I.H.P., AT e = 216 B.H.P., ij m = 67.1%. 

Heating value of gas, 139.7 B.T.U. per cu.ft., gas consumption per B.H.P. 
hour, 71.8 cu.ft., >ji = 37.76%, tj«, = 25.3%. 

In this test the heat carried off by the cooling water per hour was distributed as 

follows: 

Total heat in cooling water. 667 000 B.T.U. = 100% 

f Cylinder jacket. 467 000 B.T.U. = 70% 

Carried off in I Piston. 119 000 B.T.U. = 18% 

i Exhaust valve. 81 000 B.T.U. = 12% 

(c) Finally, the same test showed that at a maximum load equal to 650 I.H.P., 
the pump work was 36 I.H.P. =5.5% of the I.H.P. of the power cylinder, the mechan¬ 
ical efficiency was tj ot = 83.8% (including pump work), or j? m =88.8% (excluding pump 

work). 

During the first tests the indicator diagrams, with an explosion pressure of barely 
340 lbs. per square inch, showed a mean effective pressure of approximately 110 lbs. 
per square inch. At maximum load the M.E.P. exceeded even 115 lbs. per square inch. 
This pressure is surprisingly high and in conjunction with high results for and rj w 
points to excellent combustion. The compression pressure, 120 lbs. per square inch, 
exceeds by but little the inertia forces of the reciprocating parts whose weight was 
about 4.55 lbs. per square inch of piston. 

6. Ganz & Co., Budapest. The attempt to obtain thermal efficiencies as high 
as possible by increasing the compression, while correct in theory, soon meets a limit 
in practice in that the increasing compression temperatures soon cause pre-ignition. 
To avoid this, various schemes of cooling the charge either by internal or external 
means have been tried. As early as 1888, Capitaine built a kerosene engine which 
took in a little water with the mixture, by means of which it was thought to be 
possible to prevent ignition, but he did not succeed in overcoming the practical 
difficulties involved. In 1894 Donat Banki resorted to the same means of raising the 
possible compression in his engines, and five years later brought out an economically 
superior high-pressure engine, using water injection. This machine has since that 
time been built by Ganz & Co., mostly for operation with gasoline, although gas has 
lately also been tried (see below). 

Considered from a purely thermal standpoint, the use of water in this manner 
affects the efficiency of internal combustion engines in an unfavorable sense, since it 
increases the specific heat of the mixture and decreases the temperature range. 1 The 


1 Compare the investigation by Schreber in Dingler’s Polyt. Journal, 1905, p. 33. 







380 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 

theoretical loss, however, seems to be overbalanced in the Banki engine by the gain 
due to higher compression. 

The general design of the Banki engines is shown in Figs. 544-547. The only 
additional special arrangement made necessary by the method of operation is the 
double atomizer bb, placed ahead of the inlet valve a. For details of this atomizing 



Figs. 544-547. —Banki Engine. Rated H.P. = 20 at 210 r.p.m. 

arrangement see Part II. The ratio of water to 
gasoline is controlled by setting the needle valves 
of the atomizer by hand. 1 During the down 
stroke the piston along with the air also takes 
in fixed quantities of gasoline and water. The 
latter vaporizes largely during the compression stroke 
... onl y an d> in doing so, takes up so much heat 

that it becomes possible to compress to from 180 to 240 lbs. per square inch without 
pre-ignition. At the end of the up stroke the mixture is ignited by an open hot 
tube c the pressure attained being from 500 to 550 lbs. per square inch. Expansion 
and exhaust follow as usual in 4-cycle engines. 

The exhaust valve d is operated by the eccentric gear e through the wiper cam /. 
The speed is controlled by holding open the exhaust valve. This is accomplished 
through the reach rod h i, the position of which is controlled by a shaft governor in 


1 German patent No. 77764. 






























































































STATIONARY ENGINES 


381 




the wheel g. At the same time the 
spring of the inlet valve a is put in 
strong tension so as to prevent the 
opening of this valve during the miss- 
strokes. This hit-and-miss method of 
regulating the speed seems to be about 
the only one well adapted to the scheme 
•of operation, since it would be difficult to satisfactorily proportion the quantities of 
gasoline and water to variations in load. But even in this case the author s own 


Figs. 5-18 and 549.—Bdnki Gasoline Engine. Rated H.P. = 50 at 120 r.p.m. 


























































































































































































382 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


observations have shown that the cylinder temperature changes involved in the method 
of regulation may easily cause trouble. 

Operating Results, (a) Tests made by Prof. E. Meyer, 1899, on a 20-H.P. engine, 
D= 9.85", S = 15.70", compression space, .0785 cu.ft. The gasoline used had a specific 
gravity of .70 and a lower heating value of 18520 B.T.U. per pound. 1 


Table 70 


Test No. 

1 

2 

3 

4 

5 



R.p.m. 

210.9 

24.9 

.541 

71.7 

5.34 

40.0 

22.4 

25.3 

211.5 

19.3 
.590 

41.6 

3.93 

56.3 
26.0 
23.2 

212.4 

13.1 

.634 

34.8 

4.20 

72.8 

31.8 
21.6 

214.4 

6.67 

.850 

18.15 

3.21 

86.2 

28.8 

16.1 

216.2 

0 

3.26 per hr. 
11.22 

3.45 

B.H.P. 

Gasoline per B.H.P. hour.lbs. 

Injection water per hour.lbs. 

_ . water injected 

Ratio ... 

gasoline 

Cooling water per B.H.P. hour.lbs. 

Per cent heat carried off in cooling water. 

Economic efficiency.% 


The mean effective pressure, assuming ij m = .85, was computed to be 108 lbs. per 
square inch, which is rather high for an explosion engine. The indicator cards show 
explosion pressures up to 540 lbs., with compressions at 233 lbs. per square inch. In 
the consumption figures above quoted, the gasoline required by the heating lamp for 
the hot tube, which amounted to .418 lbs. per hour, is not taken into account. 

(b) Tests by Professor Jonas, with the assistance of Mr. Taborsky, made on the 
same engine near the end of 1899. The gasoline had a specific gravity of .7298 and 
a lower heating value of 18 322 B.T.U. per pound. 


Table 71 


Test No. 


1 

2 

3 

4 

5 



209.13 

209.67 

209.83 

210.50 

210.7 

B.H.P. 


26.00 

20.40 

14.85 

8.10 

0 

Gasoline per B.H.P. hour. 

.lbs. 

.493 

.523 

.582 

.727 

3.4 per hr. 

Injection water per hour. 


62.20 

35.3 

24.40 

13.72 

10.20 

_ . water injected 

Ratio ... 

gasoline 


4.84 

3.30 

2.82 

2.33 

3.00 

Cooling water per B.H.P. hour. 


29.80 

45.50 

37.70 

51.80 


Heat lost in cooling water, per cent of total 


21.7 

26.7 

23.6 

27.6 


Temperature exhaust gas. 

.°F. 

383 

383 

366 

340 

232 

Economic efficiency. 

•% 

28.0 

26.4 

23.8 

19.0 



The fuel consumption in this case is even a little better than it was in Professor 
Meyer’s tests. The effect of the water injected is quite strongly manifest in the low 
temperature of the exhaust gases. The explosion pressure varied between 555 and 
654 lbs. per square inch, depending upon the load, with compression at 234 lbs. per 
square inch. 


1 Z. d. V. D. I., 1900, p. 1062. 















































STATIONARY ENGINES 


383 


nrinril Th ’ T*** "T ‘° with «“ the same 

principle. The development apparently extended over several years and has shown 

results as indicated below. It is, however, still a question whether the Banki gas 
engme has attained marketable status. The firm itself is as yet a licensee of one of 
the builders of German gas engines. 

p / C) Th p.'i Utl ! 0r h ,? received the following tests of a 16-H.P. gas engine from 
Professor Banki himself: 




9 

No Load 


r Ull 

occ 

Half 

Explosions per minute. 

400 

1 99 A 

256 

258 

B.H.P. 

i a o 

76 

22 

0 

1.87 

51.2 per hr. 

Injection water per hour. 


ID . o 

99 

8 .61 

Gas per B.H.P hour. r „, <■* 

1 3 ftQ 

16.55 

Injection water per B.H.P. hour .... 


io . oy 

1.355 

39 43 

16.22 
1.92 

Economic efficiency. or 



. /0 | 

06 . Tto 

27.79 



The last line seems to indicate that the heating value of the gas was in round 
numbers 560 B.T.U. per cubic foot. - 

(d) Professor Schimanek, who made the above series of tests, stated in addition 
that the above figures had up to now been obtained only in two cases and that, 
probably on account of enriching the mixture by excessive use of cylinder oil (excess 
approximately 60%), the figures show results for gas consumption that are too low. 
In five other tests, made with the piston only sparingly lubricated, Schimanek found 
on the average the following results: 


Load. B.H.P. 

16.75 


n 

Gas used per B.H.P. hour. 


15.32 

O . dO 

17.80 

u 

55.0 per hr. 


referred to gas of 560 B.T.U. per cubic foot at 32° F. and 760 mm. Hg. 

The diagram, Fig. 550, with compression at 228 lbs., shows an explosion pressure 
exceeding 780 lbs. per square inch. This value, however, seems much too high and 




apparently represents an explosion accidentally very heavy. It is also quite probable 
that of this pressure at least 70 lbs. may be ascribed to inertia in the indicator. A 
good insight into the real pressure variations during the cycle is given by Fig. 551, 
which represents a distorted diagram obtained on a large scale by means of an optical 
indicator (free from inertia). The dead center positions Tvi at the end of compression 
and T Pa at the end of expansion have of course been indicated by hand afterwards.. 
































384 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


According to this card, the explosion pressure is only about 470 lbs. when the com¬ 
pression is 220 lbs., per square inch. It is also clear from the distorted diagram that 

the explosion is by no means as rapid as is indicated in the ordinary diagram, Fig. 

550. Of course the bundle of diagrams in Fig. 551 shows nothing but explosions that 

are considerably delayed, and it may therefore be assumed that the normal explosion 
would show a more rapid rise of the line. Shortly before the' end of the compression, 
T P{ , there is a marked drop in the line which is apparently due to loss of mixture 
past the piston. 



<U M-5 


■2 a Xl o <U t 3 3 O 
•Sx o-BOo xh 

5 HI 






„ » "oh 


2 >6 


ccX 




"S £ ^ 


a s O © 0) .a 

®§"-3'St.’S^ ft 

<B | g • « 
f ® 3 S > g ® 
c ® ax ~ — 

►. Q C P —' u m 

SifilMgi 

® ® ° oJ „ « 5? °.H 
S “- D ®-o «x o ? 

•c-all 

&r-|--s sa^g 

g£g§6 
g-o e g o*; 

1=0 §| 

0 ^ o ^.S § g 

0 » • • "’S-H i *| 

£ gX x 

: SS-c-S'a||«6'3 


I e 






































































































STATIONARY ENGINES 385 

Since 1892 the firm of Ganz & Co. have also built what may be considered the 
oldest gas-hammer. It seems quite natural to apply the explosfve force of gas-air 
mixtures to the operation of hammers, pile drivers, etc., and the scheme hasten 

!**?, , but only the Ganz & Co ' hammer seems to have been able to maintain 

itself. The latter, constructed according to the patents held by Bank! and Csonka is 
made only in one size, Fig. 552-555, and has the following dimensions: 

Diameter of main cylinder. 8 67 „ 

Diameter of hammer cylinder. 5 90" 

Stroke of the main piston. ^9 qq// 

Stroke of the hammer piston. 13 75 // 

Diameter of hammer rod. 2 75 " 

Weight of hammer piston, rod and hammer. 110 lbs 

R.p.m. of crank-shaft. 290 

Number of hammer blows per minute. HO 

Variable kinetic energy of one stroke, from 0 to 2880 ft.-lbs. 


IV. Americaa Gas Engines 

U The We i stin ^ house Machine Co., East Pittsburg, Pa. This company has been 
building vertical gas engines since 1896, and since about 1903 has also been building 
horizontal engines. All their product has been of the 4-cycle throttle governing type. 
Their vertical engines have all been two- or three-cylinder machines with box type 
crank case enclosing the working parts and providing splash lubrication. Several more 
or less distinct types have been brought out. 

Fig. 556 shows the earliest type, while Figs. 557-560 show a later form. The engines 
are designed to operate on natural, illuminating, or producer gas. 

The two-cylinder engines have been built in sizes from 10 to 85 B.H.P. and the 
three-cylinder engines from 40 to 650 B.H.P. The horizontal engines are made in 
large sizes only, and since their advent this company no longer builds vertical engines 
larger than 360 B.H.P. 

The principal features of the verticul engine may best be studied in Figs. 558 and 
559. The exhaust valves are located in the cylinder casting at the rear of the engine 
and discharge into a common manifold. They are operated by a cam shaft inside 
the crank case. 

The inlet valves are located in the front part of the cylinder cover and are 
operated by a cam shaft mounted above the mixture inlet manifold. 

An auxiliary igniter cam shaft parallel to the inlet cam shaft is geared to it and 
serves to vary the time of ignition while running to suit variations in the quality of 
the gas. 

The centrifugal governor controls the lift of the balanced valve. The proportion 
of gas and air admitted to this valve is regulated by the hand adjustment of cocks 
(not shown) in the gas and air lines. These serve to adjust the mixture when the 
quality of the gas changes. 


1 See, for instance, Z. d. V. D. I., 1894, p. 582; Richard, Nouveaux Moteurs 4 Gas III, p. 816; 
American Machinist, 1884, Dec. 24. 











386 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION-ENGINES 


Ignition is of the hammer break type, current for the spark coil being taken from 
either a small 100-volt dynamo through resistance lamps or from a primary or storage 
battery. 

The engine is started by compressed air introduced by special valves into one of 
the cylinders, which is for the time arranged to act as an air motor. The smaller 


Fig. 556.—Westinghouse Gas Engine. 




engines are provided with a pulley to charge the compressed air tanks after the engine 
has started. Large engines are usually accompanied by a small independently driven 
air compressor. 

The following description of the Westinghouse horizontal engines is taken from 
two sources, the Report of the Committee on Gas Engines of the National Electric 
Light Association, and Power , April 28, 1908. The engine is of the double-acting 
tandem or twin-tandem type. In the latter case the cranks are 90° apart. 

Fig. 561 shows one of the latest tandem installations, that of the new plant of the 
Union Switch and Signal Co. at Pittsburg. This is the first direct connected 60-cycle 
plant to go into service in this country with single-crank double-acting engines. 

















STATIONARY ENGINES 


387 




























388 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Fig. 559.—Cross-section, Westinghouse Vertical Engine. 



Fig. 560 —Top View, Westinghouse Vertical Engine. 




























































































































































































STATIONARY ENGINES 


389 


































390 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 

Figs. 562-567 will serve to show the general features of design. In the main 
these are as follows: 

The main frame is of heavy girder box construction, the girder surrounding the 
crank pit to stiffen the frame against the transverse stresses due to the use of the 
side crank. 

The center and rear housings, see Fig. 566, are approximately cylindrical construc- 
i tions cut away at the top to give access to the center and tailrod slippers. To 
, reinforce the center housing and to distribute the strains more evenly, removable steel 
struts are placed above the center line as shown. 

The cylinders, see Fig. 562, are supported at the ends only, giving complete 
access to the exhaust valves and piping below the cylinders. In the smaller sizes of 
engine the cylinders are symmetrical one-piece castings, having a divided jacket wall to 
avoid skrinkage stresses. The opening in the jacket wall is closed by split jacket 



Fig. 562. —General Elevation, Westinghouse Horizontal Engine. 


bands with joints of flexible packing to admit of independent expansion of cylinder 
and jacket walls. The cylinders of the larger sizes are cast in halves to obtain sound 
metal at the center of the cylinder, and these halves are then joined by means of a 
ground joint and shrink links. See Fig. 567. 

< Only the main frame is rigidly anchored to its foundation. Both the center and 

4 rear housings are free to slide on their base plates, hence permitting the unrestrained 
^ expansion of the entire construction. 

Quick access to the cylinders may be had through the inlet valve openings; no 
* heads need to be removed. 

■ The pistons are cast in one piece without chaplets. The piston is forced on the 
rod with a tapering fit and held in place by a rotating nut which is turned off flush 
with the piston face after setting up. 



























































































STATIONARY ENGINES 



In let RecMrlneL, 


\ QUERN OR 


Connection 


Inlet Rocker 


re low) 


Piston Nut Frcei 


r//X Plush With Piston 


LnySH/irT Hnd 
Eccentric Drive 
For Eoth \Jrl\jes 


Fig. 563—Cross-section through Cylinder, Westinghouse Horizontal Engine. 


301 































































































392 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 




Fig. 5G4.—Flan View, Westinghouse Horizontal Engine. 






























































































































































STATIONARY ENGINES 


393 


The piston rod is in two parts, so that the latter may be removed through the 
front and rear housings. 

The above mentioned report of the National Electric Light Association describes the 
valves and valve gear as follows: 

Valves. Instead of a single governor valve controlling all inlets from a single 
point, the governing is accomplished directly at each inlet (Eig. 563), the inlet valve 
performing the function of both mixing and governing. Thus gas and air are mixed 
only at the point of entry to the cylinder, and in exact quantity required by the 
load—a condition essential for sensitive governing. The inlet valve (Fig. 563) has two 



Fig. 565. —View showing Valve Gear, Westinghouse Horizontal Engine. 


distinct motions—a fixed vertical motion due to the eccentric, and rotation of an internal 
sleeve controlled only by the governor and provided with ports registering with 
corresponding ports in the surrounding valve cage. The essential function of this 
rotating sleeve is that of a purely throttling governor which opens or closes the 
respective ports according to the load. Hence the mixture itself remains unchanged at 
any load. Furthermore, the quantity of mixture that can be affected by a back-fire 
is a comparatively small volume contained within the inlet valve sleeve. 

Valve Gear. An excellent feature of this gear is that only one eccentric is 
employed to operate both inlet and exhaust valves, instead of independent cams or 
eccentrics (Fig. 563). This eccentric motion is transmitted by pull rods to a rolling- 
cam motion similar to that used in marine work. Inlet and exhaust are, of course, 








394 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


reversed to accomplish the valve openings at proper points in the cycle. This motion 
gives the very desirable maximum power at opening, followed by rapidly increasing 
lift as the valve opens, without bringing heavy thrusts upon the eccentric. 

The lay shaft carrying these eccentrics is driven from the main shaft by spur 
and bevel gear in place of the usual spiral gearing. This is designed to avoid the 
back-lash which eventually develops with spiral gears, due to the variable torque 
encountered. With fixed eccentrics and no lost motion, it is evident that no oppor¬ 
tunity exists for change of adjustment of valves except from wear of the rolling cams, which 
is obviated by the use of generous contact areas well lubricated. Erratic wear of the 



Fig. 566. View showing Center and Rear Housing, Westinghouse Horizontal Engine. 


gear drive is prevented by the use of the ‘‘hunting tooth,” i.e., an odd number in 
each reduction. 

Exhaust Valves. Hollow water-cooled valves of either cast iron or steel are used 
in all sizes, or a combination valve with cast-iron head and steel stem. Inside the 
bore a small tube rises to the head of the valve, constituting an overflow pipe. Cold 
water ascends through the outer annulus. The internal pipe not only prevents air pockets 
being formed but also insures that the valve never becomes dry. Stems are lubricated 
about the middle of the bushing. To remove the valves it is only necessary to drop 
out the entire cage with the assistance of a rope swing. 

The governor is of the Hartung or Jahns type, and is directly driven from the 
main shaft. It does not itself act upon the system of reach rods used to operate 






STATIONARY ENGINES 


395 


the regulating valve at each inlet, but simply operates a small balanced pilot valve 
which admits oil under pressure to one or the other end of a cylinder containing a 
piston, the rod of which controls the movements of the reach rods. By means of an 
oil relay system of this type the governor itself is relieved of nearly all work. 



Fig. 597.—Cylinder Construction, large Westinghouse Horizontal Engine. 

The ignition system of this engine deserves special mention. Make-and-break 
igniters are used, in all cases at least two, and in the larger engines three, in each 
combustion chamber. Both sides of each igniter are thoroughly insulated, making a 
double ground necessary to prevent successful operation. 



Fig. 568. —Igniter Block, Westinghouse Horizontal Engine. 


Two types of make-and-break igniters are used. One of these is the well known 
mechanical knock-off type and needs no further description; the other may be termed 
electro-mechanical. The igniter block used in the latter system is shown in Figs. 568 and 
569, while Fig. 570 shows the wiring diagram. The electro-magnet shown in Fig. 568 
has a working armature which strikes the movable electrode of the igniter plug proper 
when the circuit is established. The coil is in series with the igniter electrodes and 






396 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


serves as an induction coil. The circuit is made and broken by the commutator or timer 
shown in Fig. 570. This acts exactly like the well-known commutator used in the primary 
circuit of jump-spark systems, and the point of ignition may be changed in the usual 
way by revolving the case carrying the various circuit terminals. The electro-magnet 



Fig. 569.—Igniter Block jWestinghouse Horizontal Engine. 

acts as a tell-tale to indicate whether the igniter is working. The system also includes 
a safety stop. (See Fig. 570.) Should the speed of the engine become excessive, a dog 
is thrown out from the face of the fly-wheel by centrifugal force. This strikes the 
trigger of the trip device, breaks the circuit and puts all of the ignition out of operation. 


Z3at 

A 


*Q._ 


*S_ 


TO_ 





F;g. 571.—Diagrams from Westinghouse 
Vertical Engine. 


parts the usual gravity feed system 
valve needs to be opened and closed. 


Force feed lubrication is used for the 
cylinders, packing glands and exhaust valve 
stems. The oil supply to the cylinder is so 
timed as so reach the piston at the end of the 
expansion stroke. There are thus in all cases 
two low temperature strokes available for good 
distribution. The exact time of supplying the 
oil may be controlled by adjusting the eccentrics 
driving the oil pumps. For the other engine 
is employed, but so arranged that only a single 





























STATIONARY ENGINES 


397 


Operating Results. 

Vertical Engine, (a) Test made by C. H. Robertson, 1900. Three-cvlinder engine 
rated at 125 B.H.P., cylinder diameter = 13", . stroke = 14". Heating Value of the 
natural gas used as fuel was on the average 970 B.T.U. per cubic foot. Duration of 
each trial about one hour. 

The ratio of gas to air in all six trials was 1 to 13.7. The no-load diagram, 
Fig. 571, clearly shows the throttling of the charge and marked after-burning. The 
diagram in broken line is that obtained at maximum load, Test No. 5. 


Table 72 


Test No. 

1 

2 

3 

4 

5 

6 

Load. 

I 

i 

i 

Rated. 

Maximum. 

No Load. 



264 

264 

260 

258 

256.88 

265.5 

Comp, pressure. r in ] 

lbs. 

40.5 

60.3 

70.3 

110.0 

130.0 

25.5 

Explosion pressure. . ] middle 

per 

61.7 

99.0 

159.0 

238.0 

300.0 

32.6 

Mean eff. pressure . . L cyl. . 

sq.in. 

27.7 

46.8 

60.5 

80.0 

90.0 

8.4 

I.H.P. 


53.69 

85.55 

114.69 

139.73 

164.22 

18.57 



30.81 

64.37 

90.25 

112.86 

143.17 

0 

Mechanical efficiency. 

....% 

57.38 

75.24 

78.69 

87.93 

87.0 

0 

Gas per B.H.P. hour. 

cu.ft. 

24.3 

14.65 

12.75 

11.80 

11.70 

470 per hr. 

Thermal efficiency. 

....% 

15.6 

23.8 

26.3 

25.5 

26.1 

10.2 

Economic efficiency. 

....% 

10.76 

17.96 

20.53 

22.37 

22.56 

Exhaust temperature. 

. ,°F. 

1510 

1624 

1722 

1832 



Heat lost in cooling water . 

_ 

53.6 

30.8 


37.0 

36.7 

32.11 

42.4 

29.6 

44.3 

57.6 

32.2 

Heat lost in exhaust and radiation % 



(b) Tests of Miller & Gladden, 1900. Three-cylinder engine, rated at 550 B.H.P., 
cylinder diameter 25", stroke 30". Pistons and exhaust valves water-cooled. Heating 
value of Pittsburg natural gas on the average 1000 B.T.U. per cubic foot. Duration 
of tests 1 to 3, each 5 hours, Test No. 4, 24 hours. 


Table 73 


Test No. 

1 

2 

3 

4 

Load. 


i 

Normal. 

Maximum. 



151.7 

150.0 

149.2 

146.7 

Mean effective pressure, pounds per square 

inch. .. . 

31.0 

53.8 

75.3 

82.7 

I.H.P. 


262.5 

449.3 

621.4 

676.7 

R.H.P. 


207.7 

383.8 

553.1 

605.6 

Mechanical efficiency. 


79.15 

84.72 

88.99 

89.71 

Gas per B.H.P. hour. 

. . .cu.ft. 

15.7 

11.5 

10.45 

10.05 

Thermal efficiency. 

.% 

20.51 

25.80 

27.35 

28.38 

Economic efficiency. 

.% 

16.23 

22.02 

24.30 

25.46 

Cooling water per B.H.P. hour. . . . 

-gals. 

5.62 

4.13 

3.66 

3.54 

Cooling water to cylinder and exhaust . . . . 

% total 

70.35 

72.78 

72.22 

72.32 

Cooling water to piston. 


29.66 

27.22 

27.78 

27.67 

Mean temperature rise of cooling water . . .. 

. . . . °F. 

88.2 

84.8 

79.2 

79.2 

Heat lost in cooling water. 

.% 

38.69 

34.39 

33.34 

33.60 

Heat lost in exhaust and radiation. 

.% 

40.80 

39.83 

39.33 

38.01 


\ 



























































398 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


The figures reported in these tables deserve special attention, particularly the 
high mean effective pressure of 90 lbs. oer square inch and the excellent mechanical 



Fig. 572. 


' 2C0 lb. Spring 
Max. Compress. 137 lbs. 


efficiency .of nearly 90%. The thermal efficiency also is very high. The effect of the short 
cylinder jacket manifests itself in two directions, in a favorable sense as far as the 

low consumption of cooling water is concerned, 
and unfavorably in that the exhaust tempera¬ 
tures are high. 

(c) Tests reported by J. R. Bibbins to 
the American Society of Mechanical En¬ 
gineers, December, 1906. These tests were 
made on Westinghouse three-cylinder vertical 
engines before and after they were installed 
in the plant of the Gould Coupler Co. at 
Depew, N. Y. The fuel used is producer 
gas made in Loomis-Pettibone pressure pro¬ 
ducers operating on bituminous coal. The 
load on the plant, although to some extent 
equalized by overlapping demands from vari¬ 
ous sources, fluctuates widely, 80% above or 
below the general average, and the results 
obtained in the plant should on that account 
be of special interest. 

The plant capacity is 450 K.W. Each 
of the three engines is rated at 235 B.H.P. 
on producer gas, capable of developing 260 
B.H.P. as a maximum. The cylinder diameter 
is 19", the stroke 22", while the normal 
speed is 200 r.p.m. 

The engines were tested in the shop of the company at East Pittsburg on 
natural gas. The results of these tests are given in Table 74, and graphically in 



Fig. 573. 




































































































































































STATIONARY ENGINES 


399 


Fig. 572. The term kinetic efficiency ” corresponds to what is called thermal efficiency 
throughout this book, “ brake kinetic efficiency ” corresponding to economic efficiency. 
Fig. 573 shows some typical indicator cards obtained in these tests. 


Table 74 

ENGINE EFFICIENCY 
Average of Tests 


^(Three 19X22 ' single-acting 4-cyele gas engines. Engines Nos. 741, 742, 743.) 


Load. 

l H 

Full 

l A 

Remarks. 

Brake H.P. 

325.3 

239 

121.3 

By Prony brake 

Speed r.p.m. 

198 

203 

206.3 

By counter 

Gas per hour. 

3547 

2840 

1985 

Corr. to 60° F., 30" Hg. 

Gas per B.H.P. hour. 

10.9 

11.87 

16.36 

Corr. gas—no load—1250 ft.perhr 

Heat value gas *. 

920 

920 

920 

Effect B.T.U. per cu.ft. 

B.T.U. per B.H.P. hour. 

10030 

10910 

15050 


Brake kinetic efficiency. 

25.37 

23.32 

16.9 

B.H.P. basis 

Mechanical efficiency. 

89 

87.5 

82.5 

Estimated 

Indicated kinetic efficiency. 

28.5 

26.7 

20.5 

I.H.P. basis 

Speed variation, maximum. 

4-2% 



No load speed 206.6 

Speed variation, no to full load. 

1.8% 


Speed variation, no to half load. . 


0-7% 


Per cent full rating. 

138.4% 

101.6% 

51.6% 

On producer gas 


* Pittsburg natural gas—Junker calorimeter. Engines rated 235 B.H.P. on 130 B.T.U. producer gas. 



2500 


2000 

A 

u 

’> 

Ld 

1500 

V 

W 

a. 

E- 

<1 

C3 

loao >2 

3 

E 

y. 

a 

a 

500 


0 


125 


100 


75 


50 


25 


0 


Fig. 574.—Load Curves, Plant of the Gould Coupler Co., Depew, N. Y. 


After installation, careful daily records were kept in the plant, from which a good 
idea of plant efficiency may be obtained. Table 75 shows one of the logs typical of 
those kept from day to day. 
























































400 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


Table 75 

TYPICAL LOG. GOULD COUPLER COMPANY, DEPEW, N. Y. 
Gas Power House Daily Statement for September 26, 1905 


Time. 

Producers in 

1 2 3 

Engines in 
12 3 

1 

Pumps in 

1 2 3 

^ Station 
Watt-meters, 
Output kw. 

Volts. 

Maximum 

Amperes. 

Remarks. 

A.M. 7 

L 

A 

A L 

\ A 

A 

^ A 

1459800 

230 

1000 


8 












300 

230 

2100 

8% overload 

9 












300 

230 

2100 


10 












200 

230 

2000 


11 












300 

230 

2100 


12 




V V 






300 

230 

2100 


P.M. 1 




A L 

1 






100 

230 

2000 


2 












300 

230 

2100 


3 












200 

230 

2000 


4 












300 

230 

2000 


5 












300 

230 

2000 

Full load about 1950 

6 












200 

230 

1800 

amperes 

7 












250 

230 

1700 


8 












300 

230 

1900 


9 












300 

230 

1900 


10 












300 

230 

1800 


11 












300 

230 

1900 


12 





V 






300 

230 

1900 


A.M. 1 





2 







150 

230 

1100 


2 












300 

230 

1900 


3 












350 

230 

1800 


4 












300 

230 

1800 


5 












300 

230 

1900 


6 

V 

V 



^ V 

V V V 

250 

230 

1700 



Day Engineer. 

“ Helper. 

* ‘ Producer Man . 

“ << a 

Night Engineer .... 

‘ ‘ Helper. 

1 ‘ Producer Man 

< i < < < < 


Meter Readings: 


Hanley 6.00 p.m. 1462600 

Young Previous. 1459700 

Benarak Output. 2900 


Lea Coal, 6000; Coke,- 

Lowe Rate 2.06 lbs. per kw. hour 

Runer 

Smith 

- 6.00 a.m. 1466 000 

Previous. 1462600 

Output. 3400 

Coal, 6300; Coke,- 

Rate, 1.85 lbs. per kw. hour 


Since the above observations were made the plant has been giving much better efficiency, the coal con¬ 
sumption averaging 1.8 lbs. perK.W. hour with an 85% loading factor. This corresponds to a plant efficiency 
of over 15.4% at the engine shaft, or 14% at the switchboard. Several rims averaging 1.55 lbs. per K.W. are 
recorded, equivalent to a plant efficiency of 17.7% at the shaft, or 16.3 % at the switchboard. 


The log needs no further explanation. From the daily records thus kept Table 76, 
covering a period of ten consecutive days, has been compiled. The term “ loading 
factor used represents the ratio of average twenty-four hour station load to rated 
capacity. The nature of the varying load on any one of these days is well shown 
in Fig. 574 for Sept. 25th. 


















































STATIONARY ENGINES 


401 


Table 76 
OPERATING DATA 


450 K.W. Producer Gas Power Plant 


Date, 

September, 

1905. 

Engine 

Hours 

Run. 

Per Cent Full Output. 

Load K.W. 

Station 

Loading 

Factor.* 

Fuel Used. 

Coal 

per K.W. 

Coal per 
B.H.P. 
Hour. 

Day Run, 

K.W. 

Average. 

Rated. 

20 

71 

98 J 

4850 

202 

450 

45 

11100 

2 29 


21 

65.5 

91 

5275 

220 

450 

49 

11400 

2.16 


22 

70 

971 

6550 

273 

450 

61 

12700 

1 .93 


23 

45.75 

634 

4025 

188 

450 

41.7 

8800 

2.18 


24 

Sun. 

24 

331 

2250 

187.5 

450 

41.6 

6600 

2.93f 


25 

70.5 

98 

6400 

267 

450 

59 

12000 

1.87 


26 

69.75 

97 

6300 

263 

450 

58 

12300 

1.95 


27 

70.5 

98 

6700 

279 

450 

62 

12600 

1.88 


28 

72.0 

100 

6700 

279 

450 

62 

12900 

1 .92 


29 

70.5 

98 

6600 

275 

450 

61 

12900 

1.95 


Average . . 

63 

GO 

5565 

243.4 

450 

54 

11330 

2.04 

1.44 


Westinghouse vertical three-cylinder engines—Loomis-Pettibone producers. 

* Loading factor = per cent continuous generating capacity. n 

f Includes extra coke used on Sunday for starting new fire. 


This plant offered an unusually good opportunity of observing the effect of load 
factor upon plant efficiency. The results of this investigation are given by Mr. Bibbins 
in Fig. 575. These curves were constructed from data obtained since the main tests 



Fig. 575. —Effect of Load Factor on Plant Efficiency. 


were made and show a considerable improvement over the figures then obtained. 
This was due to more careful operation, but largely also to increased loading of the 

plant. 

















































































402 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


Table 77 

HOLDER DROP TESTS 
Summary of Results 


Test No. 

A 

B 

C 

D 

E 

Duration of Run, Minutes. 

ll 

8 

10 

10 

10 

Load, per cent engine rating .. 

no load 

25.4 

45.1 

70.6 

102.2 

Brake horse-power. 


127 0 

295 5 

258 0 

c;i i r; 

Kilowatts. 


84 1 

154 3 

9 d** Z 

9^9 n 

Speed, revolutions per min.... 

158.00 

156.0 

154.0 

152.0 

150.0 

Holder drop, feet per hour . . . 

16.91 

24.96 

32.22 

39.89 

51.60 

Cubic feet per hour, 30' 60°.... 

15760.00 

23270.00 

30050.00 

37280.00 

48200.00 

Gas consumption rate: 

(Std. gas.) 





Cubic feet per B.H.P. hour . 


183.2 

133.2 

105.5 

94.25 

Cubic feet per K.W. hour .... 


276.8 

194.8 

153.1 

137.0 

Heat value of gas: a 






Effective B.T.U. per cu.ft. .. 

106.4 

106.4 

106.4 

106.4 

106 

B.T.U. per B.H.P. hour. 


19480 00 

14160 00 

11215 OO 

i nnQn nn 

B.T.U. per K.W. hour. 


29430 00 

2079f) nn 

ift9an nn 

1UUoU.uu 

i i zan nn 

Thermal efficiency brake. . .% 


13.05 

17.96 

22.68 

14:OOU . UU 

25.36 

Thermal efficiency (elect rical)% 


11.6 

16.46 

20.96 

23.42 


Remarks. 


/ Circum. holder 
\ 110.33 ft. 


Barom. = 29.26" 
f Av. temp, of gas, 
1 71.6° F. 

/ Av.gas pressure, 
1 21 in. water 

/ Correction fac- 
1 tor—0.9642 


a Av. of all tests 


Table 78 

FRACTIONAL LOAD EFFICIENCIES 
Holder Drop Tests 


Nominal Load. 


Full. 


Overload. 


Load, brake horse-power. 

Gas consumed,* cubic feet per B.H.P. hour . . . . 

Heat consumed, B.T.U. per B.H.P. hour :. 

Heat consumed,* B.T.U. per K.W. hour . ...... 

Heat consumed,* B.T.U. per I.H.P. hour. 

Thermal efficiency, per cent brake. 

Thermal efficiency, per cent elec. 

Thermal efficiency, per cent indicated. 


125.00 

190.00 

20210.00 

30530.00 

11180.00 

12.58 

11.16 

22.75 


250.00 

127.00 

13510.00 

19700.00 

10600.00 

18.84 

17.32 

24.1 


375.00 

105.6 

11240.00 

16340.00 

9050.00 

21.66 

20.9 

28.14 


500.00 

95.00 

10100.00 

14675.00 

8460.00 

25.21 

23.25 

30.1 


550.00 

92.20 

9800.00 

14300.00 

8295.00 

25.97 

23.85 

30.7 


Equivalent Coal Consumption f for Various Producer Efficiencies, Pounds per Unit Hour 


Producer efficiency: 



100%, 

brake horse-power hour . . 

1.413 


kilowatt hour. 

2.13 

80%, 

brake horse-power hour . . 

1 .766 


kilowatt hour. 

2.663 

70%, 

brake horse-power hour . . 

2.015 


kilowatt hour. 

3.040 


0.944 
1.376 
1.181 
1.720 
1.347 
1 .964 


0.785 

0 

705 

0 

.685 

1.141 

1 

025 

0 

999 

0.980 

0 

882 

0 

857 

1 .426 

1 

281 

1 

250 

1.120 

1 

006 

0 

977 

1.63 

1 

465 

1 

426 


* Assuming same coal used on test—14321 B.T.U. 

t Standard gas—106.4 B.T.U. (effective), 60° 30" Hg. 




















































































FIFTY-ONE HOUR TEST—SIX-HOUR AVERAGES. Corrected Data Only 


STATIONARY ENGINES 


a 

A 

H 

O 

H 


a 

o 

A 

(V* 

PH 

a 

> 


S 


© 

CM 

a 

£ 

£ 

A) 


Ph 


tO 

Cl 

a 

fc 

a 

a 


S 


02 

- 

D 

o 

a 

H 

tO 


© 

CM 


CM 


CM »-h CO 

<N N W CO W O 

CJ 00 (N N iO CO 

CO (N 00 N CO 

CO CO *t CO Cl 


o 

O CO 

© CO 

CD rH 
A 


CD O 

X A rH O 

X X F-* © rH 


CD 

Cl 

CD 


C2 CD N 
CD A 


O t>» X O O CO 

© O CM rH © X 

hh CD r>- ^ 

CO CO *t C2 (N 


CD 
O © 

Cl CM 

ID rH 

A 


O 

cm to © o 

*-h © X CD 


CD 


CM 


O N CO 
h CD 


CM 

rH 

O 


X 

O 

o 

o 

N- 

o o 

o 

CM 

O 


o 

CD 

O 

X 

CD 

CM 



rH 

Cl 

-f 

CM 


CM 

o 

O 

CD 

X 

Ttl 

CM 

r-H 

CD 

■CD 

O' 


r-H 

to 



X 

X 

A 


X 

CM 

o 

CD 




tO 

X N W lO 


o o 


X X CM h X © 

co ro C 2 o ‘O o 

X X A A CO CM 


o 

t'- © 

© cm 

A H 
A 


to o 

O) N o o 

© © CM © © 


CM 

tO 

CD 


© CD O 
CD A 


CD 

o 

o 

O 

*o 


o 

iD 


CM 

CD 

H 

rH 

r-H 

CM 

O 


X 

CO h O Tf 

O lO CD H 
Cl rH Cl H 


X 

X h CO Cl 
02 CD Cl O 

h rH CM rH 


X to X CM 
*-< x cd 

rH i-H rH O 


6-12 

i 

354.5 

335.3 

487.4 

487.0 

1341.0 

O 

O 

to 

CM 

463.8 

13.30 

653 1 

X 

Ah 

X 

rH 

AH 

O 

O 

rH 

67.00 

46.00 

113.5 

to 

X 

© 

rH 

117.1 

109.8 











o 

o 















X 


to 

CD 

© 

O 


X 



CD 

X 

CM 

to 

O 

O 

o 

o 

CD 

CD 

X 

Cl 

X 

X 

A 

F'- 

© 

© 

CM 

o 

CD 

rH 

to 

o 

to 

to 

X 

CM 

rH 

02 

© 

© 

© 

© 

© 


f-H 

CM 

O 

to 

AH 

to 

Ah 


rH 

02 



CD 

A 1 


o 

Cl 



X 

X 

AH 

AH 

CM 

CM 

AH 


to 





rH 









rH 























CD 

o 















CD 


X 

o 

© 

© 





CM 

rH 

AH 

X 

tO 

o 

tO 

o 

X 

X 


Cl 

AH 

© 


A 

X 

© 

r-H 







• 

• 










1 

X 

CD 

X 

X 

Cl 

X 

CD 

X 

to 

rH 

02 

tO 

X 

© 

© 

© 


CD 

X 

X 

X 

02 

X 

AH 

CD 

rH 

02 



© 

A 

H 

rH 

CM 

o 


X 

X 

AH 

AH 

X 

Cl 

AH 


to 





rH 

rH 

rH 







>*H 























tO 

© 








iD 







CD 


CD 

© 

o 

© 





CD 

CM 

C2 

X 

r- 

o 

rH 

X 

O 

X 



© 


© 

© 

rH 

to 

CM 


CM 

TfH 

CM 

rH 

o 

tO 

X 

X 

rH 

© 

© 

E- 

X 

to 

CM 

rH 

rH 

-f 

tO 

o 

r-H 

rH 

to 

to 

rH 

rH 



© 

A 

© 

© 

Cl 

© 


X 

X 

to 

to 

AH 

Cl 

AH 


CD 





rH 

rH 

rH 

rH 






rH 























to 
















AH 


Cl 

X 

o 

© 

© 

A 

© 

© 

CM 

02 

02 

CM 

rtn 

o 

rH 

r- 

O 

rH 

CD 

to 

to 

© 

rH 

I- 

© 

rH 

r-H 








• 






• 


• 

• 

1 

X 


CM 

CD 

CM 

rH 

AH 

X 

AH 

rH 

X 

to 

A 

© 

© 

Cl 

X 

CD 

to 

to 

CM 


rH 

to 

to 

rH 

02 



© 

tO 

© 

rH 

CM 

© 


X 

X 

tO 

tO 

AH 

Cl 

AH 


to 





rH 

rH 

rH 

rH 


s 

A 

CM 

a 

£ 

a 

a 


CD 

X 


© Cl © © X 

XcDO^OOf'- 
© © A A © X 
X X 'f ^ Cl Cl 


tO 

X CM 
A ^h 
A 


CD 

CM X © O 

X A © X 

X Cl © to CM 

*H rH CD 

CD 


X 

Cl 


Cl o 

X © 

Cl o 


D 

o 

HH 

a 

Ah 


12 
p 

Sh 
02 
Oh 

S 

c3 

i 

-h 

T 0) • i 

o 02 p ^ J2 

^ S « 2 

-H H 

"§ £ 3 g 

pf ^ O C 

A Sh Sh 

,5 «H ‘-H 

,Q O . • 

. > Ah Ph 


02 
02 

£, ^ 


* 

-A 

Sh 

02 

Oh 

H 

o 

o 

Hh 

02 


9 

u 

C3 


a 

H 

D 

a 

< 

o 
— ' 

A 

H 

o 

pH 

72 

A 

o 

u 

(22 

Sh 

C2 



Ah Ah 


02 

Sh 02 
*h D W 


O 




TO TO 
02 02 

H H 

cA © 

72 72 

TO TO 
C A 
O O 
o o 

n n. 


H 

o 

D 

H 

pH 

O 


Sh 


Dh hA 

H 

H 

C2 

02 

D 

02 

H-t 

«4H 

C2 

C2 

• 

© 

© 

D 

D 

O 

o 


A 

Sh 02 

02 Sh 

Oh a 

■H 


Sh 

02 

•P 


n 


A 
02 
. H 

CQ 

02 g 


a 

n 

02 


S- 

02 

Ah 


02 

H 


W W /-*, 

H .pH 


02 02 

2 bJO bfi 

A A 

Sh Sh 

_ 02 02 

A > > 

o < < 



02 

H 

72 

rA 


H 

H 

• ^ 

CO 

S 

72 

• pH 

Sh 

>> 

rH 

A 


^3 

s 

o 

*3 

H 

H 

X 

c3 


02 

c3 


>> 

>> 

fl 


© 

© 

PH 


D 

0) 

oT 


D 




c3 

^3 

"3 


> 

> 

> 


H> 

H 

H 


c5 

c3 

a 


02 

02 

02 


KKK 


A A 
c h 

H 

pH 

• pH 

A 

• f-H 
£ 

oT 

A 
> ■ < 

A 
> 


403 


* Rate of gasification per square foot of fuel bed area of producers. 
Calorific values a’l reduced to effective at 62° F. 30-inch Hg. 



























































































404 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


Horizontal Engine, (a) Tests made by Messers Alden and Bibbins on a 23^X 
33", 500-B.H.P. horizontal engine direct-connected to a 300 K.W. generator in the 
plant of the Norton Company at Worcester, Mass., during June, 1907, and reported 
to the American Society of Mechanical Engineers, December, 1907. 

The fuel used was producer gas made in Loomis-Pettibone pressure producers. 

The coal fired in regular operation was Clearfield bituminous, nine samples of which 
averaged: Volatile matter, 19.87%; fixed C., 73.71; moisture, .87; ash, 5.54; B.T.U. 
per pound dry, 14 450 B.T.U.; as received, 14 321. The composition of the hard coal 
used to start fires was: Volatile matter, 5.20%; fixed C., 78.95; moisture 3.20; ash, 
12.65; B.T.U. per pound dry coal, 12 709; B.T.U. as received, 12 320. The average 
sulphur in the Clearfield coal was .83%. 

The tests made were of two kinds, first a 51-hour continuous test to determine 
the coal, water and oil consumption, and the mechanical efficiency; and second, a 

series of holder drop tests to determine the net heat consumption at various loads. 
It should be noted that the 51-hour test was a service run under the regular operating 
conditions with the plant in charge of the regular force. 

The main result of the holder drop tests are presented in the upper part cf 

Table 77. These results were plotted, and from these curves were obtained the figures 
in the lower part of the table. The same curves, by interpolation, also gave the 
figures in Table 78. The reduction from electrical output to brake horse-power was 
made by aid of the generator efficiency curve. 


Table 80 


FIFTY-ONE HOUR TEST 


Summary op Results (Norton Co., June 24-26, 1907) 



Load, 

Kilowatt. 

Water, 
Cubic Feet. 

Oil, 

Gallons. 

Coal,* 

Pounds. 

Quantity at finish. 

363550.0 
345710.0 
16840.0 
+ 117.3 
16957.3 

51 hrs. 
332.5 

94900.0 

63560.0 

31340.0 

2.875 

23775 

23775 

Quantity at start. 

Difference. 

2.875 

Correction. 

Corrected difference. 

31340.0 

50 hrs. 
626.8 

2.875 

48 hrs. 
0.06 

23775 

51 hrs. 

466 

Elapsed time. 

Rate per hour. 



Water, 
Cubic Feet. 

Water, 

Gallons. 

Oil, 

Gallons. 

Coal, 

Pounds. 

Rate per K.W. hour.(332.5 K.W.) 

Rate per B.H.P. hour.(482 9 B H P ) 

Rate per I.H.P. hour.(579.0 I.H.P.) 

• 

1.885 

1.3 

1.078 

14.12 

9.74 

8.075 

0.00018 

0.000125 

0.000104 

1.402 

0.965 

0.805 


1C *£ learfie . ld run-of-mine—14 321 B.T.U. per pound as fired. Average thermal efficiency of plant 
18.43%; engine, 24.93%; producer, 73.81%. Average gasification rate, 13.36 lbs. per sq.ft, per hour. 


For the 51-hour test Table 79 gives the consecutive 6-hour averages, while Table 
80 shows the total average readings. The load during this test was fairly uniform and 
as near rating as the demand of the mill would permit. During the night shifts it 
was often possible to exceed rating, thus on the night of June 24 the average load 








































STATIONARY ENGINES 


405 


was 522 B.H.P. for six hours, which corresponds to an overload of nearly 20% on the 

onffuifl' B T II 6 5ame r 6 ‘ hC ^ ating ValUe 0f the gas was below the average, 
only 109 B.T.U. per cu.ft. Fig. 576 shows 

a typical set of cards taken during the long 

run. This is one of the sets taken once an 

hour, together with the electrical readings, to 

determine the mechanical efficiency of the 

engine. From an investigation of these cards 

it appears that the mechanical efficiency at 

the average load carried was 83.5%, and at 

full load, 83.8%. These figures include all 

pump work. 

Additional determinations on oil and water 
consumption were summarized by the observers 
as follows: 

Average water consumption, for engine only, 

9.74 gallons per B.H.P. hour with 66° F. inlet 
temperature and 47.1° F. rise. 

Average cylinder oil consumption, 1.44gallons 
per 24 hours, equivalent to 0.6 gallon per operating 
day, or 3.2 gallons per operating week. 

Average producer efficiency, 74.4% at full 
load, 73.8% at average test load—both based 
upon lower or effective heat value of gas. 

Producer gas, average, 114.6 effective 
B.T.U. during 51-hour test; maximum vari¬ 
ation 11.5% above and below mean. Differ¬ 
ence between total effective heat values— 
about 4f%. 

This important test, complete in all respects, 
was supplemented by speed variation tests, the 
results of which are given in Table 81. 



/7vr./yr.^:^sf47 

Lo^o/^yy.-34-o 




Fig. 576. 


Table 81 

SPEED VARIATION TESTS 


Speed, revolutions per minute. 

155 

154.0 

152.0 

150.0 

149.0 

148.0 

Volts. 

255 

255.0 

257.0 

258.0 

258.0 

257.0 

Amperes. 


327.5 

665.0 

955.0 

1303.0 

1347.0 

K.W. 


86.1 

170.8 

246.6 

336.1 

346.0 

B.H.P. 


129.6 

247.6 

356.5 

489.3 

503.0 

Per cent full rating. 


25.9 

49.5 

71.2 

97.9 

100.5 

Speed drop, per cent ± mean. 


0.819 

0.958 

1.597 

1.916 

2.236 


Instantaneous Load Test June 27, 1907, 6 p.m. 

No load to full load, 280 volts, 1190 amperes, 345 kilowatts, 502 brake horse-power 


No-load speed.155 revolutions per minute 

Load thrown on.148 revolutions per minute 

Load thrown off.155 revolutions per minute 

Difference. 7 revolutions per minute 

Speed variation.4.6% of total —2.3% imean speed 








































406 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


Other data available on the operation of this plant is given in Table 82, which 
was published in Power for July, 1907. This table gives the average operating results 
for seven consecutive weeks beginning with Jan. 6, 1907. The average coal consump¬ 
tion per B.H.P. or per K.W. hour for this entire period is some 25% higher than in 
the 51-hour test made by Messrs. Alden and Bibbins in June of the same year. This 
is largely due to the fact that the loading factor during this time was only about 
three-fourths rating, while during the June tests it was practically unity. A slightly 
lower grade of coal may also have affected the results to a small extent. 


Table 82 

SUMMARY OF OPERATING DATA OF GAS-POWER PLANT OF THE NORTON COMPANY, 

WORCESTER, MASS. 


"Week Ending. 

Hours Run. 

Output, 
K.W. Hours. 

Coal Burned, lbs 

.* 

Coal per K.W. Hour. 

Producer. 

New Fires. 

Total. 

Producer. 

Total. 

January G . 

45-5 

10150 

20690 

1400 

22090 

2.03 

2.17 

January 13. 

50-5 

11900 

24600 

2700 

27300 

2.07 

2.29 

January 20. 

55 

13200 

24530 

2570 

27100 

1.86 

2.05 

January 27. 

54-40 

12400 

22500 

2000 

24500 

1.84 

2.00 

February 3 . .. . 

55 

13600 

23S00 

2400 

26200 

1.75 

1.93 

February 10 ... . 

55 

13000 

24265 

3060 

27325 

1.87 

2.10 

February 17 .... 

55 

13000 

23700 

|2500 

26200 

1.82 

2.01 

Total. 

369-50 

87250 

164085 

16630 

180715 



Average. 

53 

12464 

23441 

2376 

25816 

1.88 

2.07 


Week Ending. 

Generator 

Efficiency. 

Coal per B.H.P. Hour. 

Heat Consumption.t 

Plant Efficiency.! 

Station 

Loading 

Factor. 

Average 
Load 
per Cent 
Engine 
Rating. 

Producer. 

Total. 

B.T.U. per 
K.W. Hour. 

B.T.U. per 
B.H.P. 
Hour. 

Electrical. 

Brake. 

January 6.... 

92 

1.39 

1.49 

28400 

19500 

12.00 

13.1 

20.2 

75.0 

January 13 ... . 

92 

1.42 

1.57 

29000 

19900 

11.75 

12.9 

23.6 

79.0 

January 20 ... . 

92 

1.28 

1.41 

26000 

17900 

13.10 

14.2 

26.2 

80.0 

January 27 ... . 

92 

1.26 

1.37 

25800 

17600 

13.20 

14.5 

24.3 

74.8 

February 3.. . 

92 

1.20 

1.32 

24500 

16800 

13.90 

15.2 

27.0 

82.4 

February 10 . . . 

92 

1.28 

1.44 

26200 

17900 

13.00 

14.2 

25.8 

78.8 

February 17 . . . 

92 

1.25 

1.38 

25500 

17500 

13.40 

14.5 

25.8 

78.8 

Total. 










Average. 

92 

1.29 

1.42 

26300 

18050 

12.96 

14.1 

21.9 

78.7 


* Med. grade Penn, bituminous coal up to January. 21. Pocahontas slack thereafter. Assumed, 14000 
B.T.U. average. 

t For continuous running, not including 10% loss in pulling fire weekly. 

J Estimated from previous runs. 


















































STATIONARY ENGINES 


407 


(&) Tests made under the direction of Mr. W. Dalton, Chief Engineer, on a 
horizontal tandem gas engine in the plant of the American Locomotive Co. at Rich¬ 
mond, Va., and reported by J. R. Bibbins to the Convention of the American Insti¬ 
tution of Electrical Engineers, June, 1908. 

The engine is of the usual double-acting tandem construction, with cylinder 
diameter equal to 23^" and stroke equal to 33". The fuel used in the producers was 
bituminous coal. The producer is an R. D. Wood induced draft bituminous type 
with water-sealed bottom and centrifugal type tar extractor. An internal belt evapor¬ 
ator located above the fire bed furnishes the vapor required for dissociation. Wet and 
dry scrubbers are also used in addition to the centrifugal tar extractor. A large 
gasometer serves as a mixing chamber and provides ample storage for a short period 
of producer shut down, as well as for starting up the entire plant. 

The demand on the plant is fairly constant owing to the overlapping power 
requirements of a considerable number of electrically driven machines throughout the 
works. The test was continued for practically four weeks, the first half of this time, 
223 hours, at nearly full load, and the remaining two weeks on f load and \ load. 
The observations were made entirely by the staff of the locomotive company in order 
to determine the fulfilment of guarantees. 

The principal results of the test are shown in the following table: 

Table 83 


GENERAL RESULTS OF TEST 



Full Load. 

f Load. 

i Load. 

Length of run, hours. 

223 

125 

136 

Average load, kilowatts. 

312.3 

228.3 

159.6 

Average load, computed B.H.P. 

455.0 

333.0 

238.0 

Load, per cent engine rating. 

91.0 

67.6 

47.5 

Load, per cent generator rating. 

104.0 

77.2 

53.2 

Coal gasified, lbs.. 

115289 

54143 

47775 

Coal gasified per hour ... 

517.0 

433.0 

351.0 

Output kilowatt hour. 

69650 

28540 

21710 

Pounds coal per kilowatt hour. 

1.654 

1.697 

2.20 

Pounds coal per kilowatt hour guaranteed. 

1.93 

2.10 

2.64 

Pounds coal per B.H.P. hour. 

1.14 

1.31 

1.56 

Average heat value of eoal R T\TT. 

14392 

14392 

14392 

p T XT per kilowatt hour , T T r . 

23700 

27280 

31650 

B.T.U. per B H P. hour.... . 

16415 

18710 

21670 

Per cent thermal efficiencv brake. . 

15.51 

13.6 

11.75 

Per cent thermal efficiency, electric. 

14.35 

12.65 

10.78 






Coal .—Pocahontas run-of-mine; average heat value dry sample, 14 703, as fired, 14 392; volatile matter, 
22.8%, ash, 4.5%, sulphur, 1%. 

Test. —Aug. 12, 7 a.m. to Sept. 7, 12 m. 


At a load of 312 K.W. the engine is running slightly under rated load, and, 
on the basis of the figures observed, it is estimated that the coal consumption at full 
load, 350 K.W., would have been 1.59 lbs. per K.W., or 1.09 lbs. per B.H.P. 

During the four weeks test there were of course several periods of idleness, as the 
plant is not operated over Sunday. The producer operator estimates that from 6 p.m. 
on Saturday to 7 a.m. on Monday 1700 lbs. of coal on the average were required to 
cover stand-by losses. 
























408 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


Determinations of oil and water consumption were also made during the 223-hour 
test. The average consumption of cylinder oil was .09 gallons per hour, equivalent 
to 4.9 gallons per week for a 10-hour working day. The average consumption of 
cooling water was 6.56 gallons per B.H.P. hour at heavy load, the inlet temperature 
varying from 75-80° F., and the outlet from 140-150°. 

2. De La Vergne Machine Co., New York. This firm builds four distinct types 
of engines, the Korting 2- and 4-cycle engines, which they now call type KT and type 
KF, the Hornsby-Akroyd 4-cycle horizontal oil engine (type HA) and the De La 
Yergne 2-cycle vertical oil engine (type S). (Two other oil engines, to be known as 
types FH and FV, horizontal and vertical respectively, have also been produced in which 
heavy crude and fuel oils are completely burnt with the assistance of highly com¬ 
pressed air, but these two types are at this time (Dec., 1908) only emerging from the 
experimental stage. They are omitted in this edition.) 

The first two of the engines above mentioned have already been described in some 
detail on p. 334 to 341, and hence it will only be necessary to point out the differences 
in the American types where they exist. 


Table 84 


Engine. 


Single. 


Twin. 


Single. 


Twin. 


B.H.P. with producer and blast-furnace gas 
B.H.P. with natural gas. 


500 

600 


1000 1000 
1200 1200 


2000 

2400 


Power cylinder. 

Revolutions per minute. 

Piston speed, feet per minute 

Main bearings. 

Outboard bearing. 

Main crank pin. 


Main cross-head pins. 

Main cross-head shoe bearing surface. 

Main piston rod, diameter. 

“Light” fly-wheel, factory work,diameter. 

“ ‘ “ , “ “ , weight,. 

“Heavy” “ .electric “ , diameter. 

“ “ , “ “ , weight. 

Bore of fly-wheels. 

Diameter pumps for furnace gas. 

Diameter air pump for producer and natural gas 
Diameter gas pump for producer and natural gas 
Stroke of all pumps. 

Diameter gas inlet pipe .. 

Diameter exhaust pipe. 

Diameter cooling water pipe. 

Length over all (heavy fly-wheel). 

Width over all *. 

Shaft center to floor level. 

Shipping weight without wheel. 


25*"X45" 

38"X60" 

100 

70 

750 

700 

12f"X23f" 

18"X33" 

12"X21" | .... 

14X25i"i .... 

12f"X13i" 

18"X19i" 

two | four 

two | four 

H® 

X 

00 

12J"X11J" 

19f"X34f" 

27*"X52" 

6* 

10 

17'. 6 

22' 

36000 

66000 

17' 10'' 

23' 

72000 

130000 

16'' 

23f" 

28f" 

381" 

28£" 

381" 

251" 

34f" 

33£" 

49*" 

13" 

18" 

17" 

24" 

3i" 


39' 4" 

55' 11" 

23'5" | 33'6" 

30' | 45' 

34"+ 6" 

39"+ 7" 

132000 | 250000 

380000 | 746000 


Including extended shaft, or generator between the cylinders of a twin. 


















































STATIONARY ENGINES 


409 


Table 85 

PRINCIPAL DIMENSIONS OF 4-CYCLE SINGLE-ACTING GAS ENGINES (Ivorting Patents) 

Single-cylinder Engines 


Rated B.H.P with producer gas, 82% 
and 68 lbs. M.E.P. 

75 

85 

100 

125 

150 

175 

Do. with natural or illuminating gas,* 
78 lbs. M.E.P. 

85 

100 

115 

145 

175 

200 

Revolutions per minute, Std. 

170 

160 

155 

145 

135 

130 

Diam. and stroke of motor cylinder, 
in millimeters. 

430X700 

460X755 

490X800 

540X876 

590X955 

650 X 955 

Do., in inches. 

16*X27* 

18£X29* 

19£X31£ 

21£X34£ 

23£X371 

251X371 

Piston speed, feet per minute. 

780.8 

792.6 

813.6 

833.5 

846.0 

814.6 

Crank pin, diameter X length. 

7|X8R 

8*X9* 

9*X10£ 

10*X11 

11X12* 

121X14* 

Wrist or piston pin (bearing, diam¬ 
eter X length . 

5 A X8£ 

5MXSR 

OiX 9* 

6RX101 

7£X 11 

9*X 121 

Main bearings (2), diam.Xlength. 

7RX14* 

S£X16* 

8iXl7ft 

9|X18* 

101X191 

n*x2i! 

Outboard bearing (1), diam.Xlength. 

6£X12* 
10' 8" 

6iXl2* 
11' 3" 

6£X12* 

9X27£ 

94X23 
13' 54" 

104X314 

14' 

Diameter of heavy fly-wheel. 

12' 

12' 7£" 

Diameter of light fly-wheel. 

10'11" 

10' 10" 

12' 44" 

12' 6" 

13' 5£" 

13' 9" 

Bore of (either) wheel or of pulley.... 

9* 

10* 

10f 

HI 

121 

14* 

Weight of heavy fly-wheel,t net lbs. . 

16700 

17200 

21000 

27200 

40600 

50200 

Weight of light fly-wheel, net lbs . . . . 

9900 

11400 

13200 

17600 

24000 

30000 

Weight of engine -without wheel,gross 

24000 

28800 

34600 

42800 

48200 

50300 

Weight of engine without wheel, net . 

22500 

26500 

31500 

39100 

43900 

45600 

Driving pulley (split), diameter. 

4' 6" 

5' 

6' 

7' 

9' 

10' 

Driving pulley width of double belt, in. 

12 

14 

14 

16 

16 

18 

Length over all incl. heavy wheel. 

17' 

18' 8" 

19' 

21' 

22' 4" 

23' 7" 

Width over all incl. outboard bearing. 

9' 2" 

11' 9" 

12' 1" 

11' 9" 

12' 4" 

13' 10" 

Height, bottom of frame to top of 
heavy wheel. 

7' 8" 

8' 3" 

8' 8" ‘ 

9' 3" 

9' 9" 

10' 2" 

Height, bottom of frame to center of 
crank-shaft. 

25f" 

26*" 

28f" 

28f" 

28"! 

31£" 

Weight of frame casting, rough, lbs. 

8750 

10000 

10780 

12175 

16075 

23340 

Size of air pipe, inches. 

7 

7 

7 

8 

8 

10 

Size of exhaust pipe, inches. 

7 

7 

8 

8 

9 

12 

Size of producer gas pipe, inches 

7 

7 

7 

7 

7 

8 

Size of natural or ilium, gas pipe, in. 

4 

4 

4 

5 

5 

6 


* With blast-furnace gas, M.E.P. = 60 lbs. per sq.in. and B.H.P. is but 88.2% that with producer gas. 
t Light fly-wheels are for general power engines, pumping, etc. Heavy wheels are for D.C. direct-con¬ 
nected or belt-driven generators of any kind. Direct-connected A.C. generators require twin engines and 
extra heavy fly-wheels. All wheels made in halves. 


Table 86 

TWIN-CYLINDER ENGINES 


Rated B.H.P. with producer gas,68 lbs. 

150 

170 

200 

250 

300 

350 

Do. with natural or illuminating gas, 
78 lbs. M.E.P. 

170 

200 

230 

290 

350 

400 

Rev. per min. (same as single cyl.) . .. 

170 

160 

155 

145 

135 

130 

Size of cyls. (same as single cyl.), mm. 

430X700 

460X755 

490X800 

540X876 

590X955 

650 X 955 

Width of engine,* including a genera¬ 
tor placed at end of crank-shaft.... 

17' 

19' 

21 ' 

22 ' 6" 

23' 6" 

25' 

Do. excluding the above generator. . 

13' 

14' 

15' 

16' 

17' 3" 

18' 6" 

Do. with generator or pulley in middle 

15' 9" 

17' 8" 

18' 2" 

18' 4" 

19' 8" 

22 ' 1" 

Weight of whole engine without fly- 
wheel or pulley, gross lbs. 

46500 

55800 

66600 

83000 

93000 

97500 

Do. net lbs. 

43600 

51400 

61000 

75800 

85200 

88500 

Driving pulley (split), diameter. 

6 ' 

7' 

7' 6" 

9' 

10 ' 

28" 

12 ' 

28" 

Driving pulley, width of double belt, in. 

18" 

20 " 

22 " 



* In each case is the fly-wheel assumed as between the two cylinders. 

Note. —Fly-wheel and other parts are alike for twin- and single-cylinder engmes. 


















































410 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


1 . The Korting 2-cycle engine. (Type KT.) The main features of the American 
design of this engine are shown in Fig. 577. This picture, however, does not show the 
latest governing gear which is illustrated in detail in Figs. 578 and 579. The former 
indicates that the air pump and its valves operate in the ordinary way. The gas pump 
cylinder, however, is provided with a ring of ports at the middle so that for one-half of 
the discharge stroke the piston forces the gas back into the suction main. Each end of 
the gas pump is fitted at the bottom with suction and discharge valve and at the 
top with a by-pass valve. This valve is positively operated by eccentric and reach 
rods over the cylinder, as shown in Fig. 578. It is tripped by a gear similar to a 
Corliss gear, Fig. 579, at a point in the second half of the gas pump discharge stroke 
determined by the governor. The method of operation of this gear becomes clear from 
a study of the figure. As long as this valve remains open, the gas continues to be 
forced back into the suction main. As soon as the gear trips, however, the spring 
shown closes the valve and the gas is forced through the discharge valve at the 
bottom of the pump and into the power cylinder for the remainder of the pump 
stroke. 

The table on p. 408 shows the principal dimensions of De La Vergne-Korting 2-cycle 
gas engines. 

2. The Korting 4-cycle Engine. (Type KF.) This engine has already been 
described (p. 334) and there is nothing radically different in the American design. 
The tables on page 409 give the principal dimensions of the various sizes made by the 
De La Yergne Co., together with some other data. 

The features of the De La Yergne suction producer used with these 4-cycle engines 
are in general those of the Korting suction producer already described on p. 274. 

The following table gives the principal dimensions required for electric power 
plants equipped with De La Yergne type KF engines and prod cers. The letters refer 
to Figs. 580 and 581, which show two lay-outs, one for direct connected, the other 
for belted units. 


Table 87 


ELECTRIC POWER PLANTS WITH DE LA YERGNE GAS ENGINES AND SUCTION GAS 

PRODUCERS 


Arranged for Single-cylinder Units of 75 to 175 B.H.P. Capacity 


A. Gas engines (Korting type). 

B. Gas producers. 

C. Vaporizers. 

D. Coke scrubbers. 

E. Sawdust purifiers. 

F. Expansion tanks. 


G. Electric motor, for 

H. Belt-driven air compressor. 

I. Compressed air receiver. 

J. Electric motor, for 

K. Blower for producers. 

L. Exhaust silencers. 


TABLE OF APPROXIMATE DIMENSIONS IN FEET 


Size 

Direct Connected (Fig. 580). 

Belted (Fig. 581). 

B.H.P. 

75 

85 

100 

125 

150 

175 

75 

85 

100 

125 

150 

175 

M 

22 

23 

24 

25.5 

27 

28.5 

15 

16 

17 

17 

18 

19 

N 

31 

33 

34 

35 

37 

40 

38 

38 

40 

41 

42 

43 

0 

17 

17 

18 

18 

19 

19 

15 

15 

15 

15 

16 

16 

P 

20 

21 

22 

23.5 

25 

26.5 

22 

23 

24 

25.5 

27 

28 

Q 

24 

24 

28 

29 

32 

32 

13 

13 

15 

15 

18 

18 

R 

15 

15 

15 

15.5 

16 

16 







_ 

































STATIONARY ENGINES 


411 

















ARRANGEMENT OF GAS-PUMP AT/ON 


412 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


<0 




Fig. 578. 
























































































VAl VE GEAR MOT/ONro* GAS-PUrtP 

C> J £* 61 /VC 


STATIONARY ENGINES 


413 


s 





Fig. 579. 


















































































































































414 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Producer Room 









































































































































































STATIONARY ENGINES 


415 




i. 581, 




















































































































































































416 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 

3. The Hornsby-Akroyd Oil Engine. (Type HA). This is a very popular English 
4-cycle oil engine, placed on the American market in 1895. Fig. 582 is a general 



Fig. 582. 



vapor' i 


. i i ~r~' 


i " i..i...,, .r... ::::j 


ssmmSm m 






view of a 20 B.H.P. size of this type. The constructive details are illustrated in 
Figs. 583, 584 and 585, from which its operation should be clear without much explana- 





























































417 


STATIONARY ENGINES 



tion.- On the suction stroke of the piston the oil is injected by a plunger pump 
operated by the am valve cam on the lay shaft, through an atomizing plug into 2 

vaporizer chamber which is a cast-iron vessel 
connected to the cylinder by a narrow neck. 
The oil vaporizes when coming in contact with 
the hot walls and the mixture is formed when 
the piston on the return stroke compresses the 
air into the vaporizer. The compression is such 
that the mixture ignites when the piston reaches 
the dead center, no igniter of any kind being 
required. 

The vaporizer is heated before starting by 
means of a blow lamp. In order to accomplish 
this inside of ten minutes the De La Vergne 
Co. make a special heating lamp which works 
on the principal of the ordinary gasoline torch, 


Fig. 584. 


Fig. 585. 



Fig. 586. 


Fig 587. 


but uses kerosene as fuel. The lamps used for the smaller size of engine are shown in Fig. 
586, comprising one or more burners, while Fig. 587 illustrates a special air compressor 











































418 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


outfit and lamp used for the larger sizes. The compressor and receiver can also be used 
for starting the engine. 

After starting, the heat of compression and explosion is sufficient to maintain the 

vaporizer cap at a dull red heat. A protecting hood 
placed over the cap at starting can be lifted off if the 
engine tends to get too warm by running continuously 
at full load. Using moderate pressures it is practically 
impossible to burn kerosene or crude oil without the 
formation of some carbon deposit. Recognizing this 
fact, the oil is in this engine not injected into the 
cylinder where the deposits might cause trouble, but 
into a separate vaporizing chamber. 

Originally the vaporizer consisted of a single 
gun-iron casting, but about ten years ago it was 
found that the compression could be increased without 
incurring pre-ignitions, by simply water-jacketing about 
one-half of the surface of the vaporizer. In this way the 
economy was improved and the horse-power increased 
by 20% without adding to the number of revolutions. 

Having the vaporizer in two parts, as indicated 
in Fig. 583, the removal of the deposits becomes a 
very simple matter. In the large sizes the hot cap has a hand-hole plate held 
in place by yoke and set screw. Joints are made by means of an endless copper wire 



ring recessed in the water-jacketed part of the vaporizer. How often the operation 
of scraping out the carbon may have to be undertaken depends altogether upon the 



Fig. 588. 























































































































STATIONARY ENGINES 


419 


nature of oil used, and can only be determined by trial. At best it will have to be 
done once a month, at w r orst, once a day. Unless very excessive, the deposits do 
not interfere with ordinary operation and by keeping duplicate caps on hand an 
engine can be operated for at least 23+ hours each day in the worst cases. 

The speed of Hornsby-Akrcyd engines is regulated by proportioning the quantity 
of oil to the load. The fly-ball governor is shown in Fig. 582, and in greater detail 
in Fig. 588. It acts upon a by-pass valve so that the oil discharged by the pump 
(which is always a constant quantity depending upon the setting of the effective 
pump stroke) is divided into two parts, one going to the vaporizer, the other returning 
to the fuel tank in the engine base. 

These engines are made as single-cylinder units in sizes from 10 to 125 B.H.P., 
and as twin-cylinder units in sizes from 40 to 250 B.H.P. Figs. 589, 590, 591 and 592, 
together with the tw r o tables following, show the principal dimensions. 


Table 88 

PARTICULARS OF HORNSBY-AKROYD OIL ENGINES —(Horizontal) 

Single-Cylinder 


Brake H.P. 


10 

15 

20 

25 

35 

50 

85 

125 

Diameter and stroke. 


9.5X12 

11X15 

13.5X16 

14.5X17 

16X20 

18.5X25 

23X28 

27.5X33 

Revolutions per minute . . 


290 

260 

220 

220 

220 

180 

180 

145 

Number of flv-wheels .... 


2 

2 

2 

2 

2 

1 

1 

1 

Diameter of fly-wheels. . . 

. .ft. 

4.5 

4.75 

5 

6 

6 

9 

11 

12 

Face. 


6 

7 

7 

7 

7.5 

10.5 

11.5 

13 

A . 


18.25 

22.62 

25.25 

28.25 

32 




B . 


18.25 

22.62 

26.5 

28.25 

32 

38 

44.5 

53.5 

C . 


4' 2.5" 

5' 0.5" 

5' 9.75" 

6' 10.5" 

7' 6.75" 

9' 3" 

10' 4.5" 

12' 7" 

L . 


7' 9" 

9' 4" 

10' 

10' 9" 

12' 3.5" 

14' 6" 

17' 4" 

20' 6" 

Floor line to shaft center . 


29.5" 

31.5" 

32" 

38" 

38" 

36" 

36" 

45" 

Approx, ship, weight, . . . . 

.lbs. 

5000 

7500 

10400 

13600 

16000 

24000 

41000 

70000 


Table 89 

PARTICULARS OF HORNSBY-AKROYD OIL ENGINES— (Horizontal) 
Twin-cylinder. With One Fly-wheel 


Brake H.P. 

Diameter and stroke. . . 
Revolutions per minute 

Diameter of wheel. 

Face of wheel. 


With shaft coupling outside . .. . 
With pulley inside. 



Floor line to shaft center. 

Approximate shipping weight . . . lbs. 


40 

13.5X16 

220 

7' 

8 " 

4' 4 " 

8' 5" 
o' 6" 

9' 1.5" 
IP 
32" 
17200 


50 

14.5X17 

220 

7' 

8 " 

4'8.75" 
9' 3" 

6 ' 3" 
10 ' 2 " 
IP 3" 
32" 
22000 


70 

16X20 

220 

7'10.5" 
9" 

5' 6" 
10 ' 6 " 
7' 

IP 5" 
13' 3" 
32" 
27000 


100 

18.5X25 

180 

9' 

10.5" 
6 ' 6 " 
12' 7' ' 
8 ' 

13' 2" 
14' 6" 
36" 
39000 


170 

23X28 

160 

IP 

11.5" 
7' 6" 
13' 3.5" 
8 ' 6 " 
14' .1" 
17' 4" 
36" 
66000 


250 

27.5X33 

145 

12 ' 

13" 

9' 

16' 9" 

IP 

17' 10" 
20 ' 6 " 
45" 
115000 


Hornsby-Akroyd engines have found extended application in this country, there being 
not less than 5315 B.H.P. in one industrial establishment alone. They have also been 
extensively used for driving the oil pumps in oil pipe line stations. An example of a 
modern plant of this kind is that of the Gulf Pipe Line Co. in Texas, the lay-out 
for which is shown in Fig. 593. The oil used for fuel is taken directly from the line. 












































420 CONSTRUCTION. ERECTION. TESTS OF INTERNAL-COMBUSTION ENGINES 



































































































































STATIONARY ENGINES 


421 



There are three stations, co\ ering a total length of about 300 miles. The pressure 
pumped against is about <00 lbs. per sq.in., each station pumping in the neighborhood 
of 576 000 gallons per day of 24 hours. 

4. The De La Vergne Vertical 2-cycle Oil Engine. (Type S.) This is a small 
engine recently developed. Fig. 594 explains the construction and at the same time 
shows the appearance of the stationary type, Fig. 595 illustrating the marine type. 

Both of these illustrations show two-cylinder 
engines, but single-cylinder machines are 
also made, as may be seen from the follow¬ 
ing table of dimensions: 


Fig. 594. 


Fig. 595. 


Table 90 


PARTICULARS OF DE LA VERGNE VERTICAL OIL ENGINES 
All Dimensions in Inches 


Brake H.P. 

No. of 
Cylinders. 

Diameter 
and Stroke. 

R.P.M. 

Floor Space 
(Size of Base). 

From Floor 
or Bottom 
of Base to 
Shaft Center. 

Height 
above Shaft 
Center. 

Diameter 
and Face of 
Standard 
Fly-wheel. 

Total 

Shipping 

Weight. 

Approx. 

4 

1 

5X 5.5 

650 

STATIONARY E 

14X22 

NGINES 

14 

23.75 

20X4.5 

450 lbs. 

8 

2 

5X 5.5 

650 

23X22 

14 

23.75 

20X4.5 

700 “ 

74 

1 

7X 7.5 

500 

19.5X29.5 

164 

32.62 

28X5.25 

1100 “ 

15 

2 

7X 7.5 

500 

30.5X29.5 

164 

32.62 

28X5.25 

1600 

124 

1 

8X10 

450 

23 X 38.25 

20 

43 

30X5.25 

1800 “ 

25 

2 

8X10 

450 

36.5X38.25 

20 

43 

30 X 5.25 

2700 “ 

4 

1 

5X 5.5 

650 

MARINE ENGINES 

6X15 1 5.625 

23.75 

20X4.5 

350 lbs. 

8 

2 


650 

14.75X15 

5.625 

23.75 

20X4.5 

550 “ 

71 

15 

1 

7 X 7.5 

500 

10X20 

7 

32.62 

26 X 5 25 

900 “ 

2 

7X 7.5 

500 

21X20 

7 

32.62 

26 X 5.25 

1300 “ 

121 

25 

1 

8X10 

450 

11 .5X26 

9.5 

43 

30X5.25 

1400 “ 

2 

8X10 

450 

25X26 

9.5 

43 

30X5.25 

2200 “ 































422 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


The engine operates on the ordinary 2-cycle principle, compressing the air for 

scavenging and charging in the crank case. The vaporizer chamber V, Fig. 594, is 

merely the head or extension of the cylinder and not a separate chamber with a con¬ 
tracted neck, as in the Hornsby. Oil is injected through the plug N by means of a 

plunger pump P just as the piston reaches the upper dead center. The oil vaporizes 
almost instantly as the nozzle . N projects into the cylinder and is thus kept hot. A 
cross-section of the spray nozzle is shown in Fig. 596. K, Fig. 594, is a plunger 



Fig. 596. 


arrangement by which oil may be injected by hand into the cylinder on starting. 
It can also be used for changing the pump stroke. 

For the purpose of keeping the vaporizer hot, these engines, like the Hornsby, 
are governed by regulating the quantity of oil required per power stroke. The 
governor is shown in Fig. 597 and its action is described as follows in the catalogue 



of the company: Frame F is fastened with two studs concentrically to the inside of 
the fly-wheel. To this frame is hinged the cam ring R which, on the fly-wheel side, 
has a projection B, and back of that, cam projection C. 

Once in each revolution of the fly-wheel, cam C pushes roller A, thus moving oil 
pump plunger P. The stroke imparted to the plunger is gauged by the lever L 
pivoted to frame F. In the middle of lever L is a wedge W, which more or less 
separates cam ring R from frame F. 
























































STATIONARY ENGINES 


423 


When the engine is started, the full charge of oil is injected, and the governor 
does not act till the normal speed is approached. Under light load, the speed tends 
to accelerate when, by the action of centrifugal tendency, the counterweight H will 
overcome the tension of spring S and thus draw the wedge away from the crank-shaft, 
more or less. Spring E always keeps roller A riding on cam ring R. If, therefore’ 
the wedge is withdrawn, buffer G moves away from plunger P, and to that extent 
the effective lift of C is reduced, and with it the amount of oil injected. If, at 
excessive speed, the wedge is entirely withdrawn, spring E can push buffer G so far 
away from plunger P that, even when pushed back again by C, the plunger will not be 
touched, and no oil will be pumoed, compelling the engine to come down to its normal 
speed. 

The two concentric slots in frame F allow of making an adjustment when it is 
desired to greatly change the normal speed of the engine, the effect being that the 
time of injection, which is coincident with the time of ignition, is advanced or retarded 
relative to the upper dead point of the crank. For smaller speed variations, the 
spring S and weight H can be shifted in the holes of lever L. By adjusting the lock 
nuts N of the throttle bar T, which limit the stroke of plunger P by means of bell 
crank M, the oil supply can also be diminished and thus the speed reduced by about 
10 %. 

Operating Results. The Korting 2-cycle Engine. There are apparently no tests 
available on Korting 2-cycle engines in America. The company bases the rating for 
these engines upon a mean effective pressure of 65 lbs. for producer and blast furnace 
gas, and 75 lbs. for natural gas. In order to prevent pre-ignition at the high com¬ 
pression carried, 150 lbs., the charge is diluted with cooled exhaust gas. The mechan¬ 
ical efficiency is from 70-75%, including the pump work. Engines are able to carry 
about 15% overload. A recent guarantee on a 500 B.H.P. engine running on natural 
gas was 11000 B.T.U. per B.H.P. hour, which corresponds to an economic efficiency 
of 23.1%. 

The Korting 4-cycle Engine. In these engines, running on producer gas, the 
compression pressure at full load is from 160 to 180 lbs. gauge, the maximum pressure 
from 300 to 350 lbs., terminal pressure about 32 lbs., M.E.P. about 65 lbs. The 
mechanical efficiency varies between 82 and 87%. 

(a) Test of 75 B.H.P. engine and producer at the Merrimac Chemical Co.’s 
Works, North Woburn, Mass., made by Professor Miller of the Massachusetts Institute 
of Technology, for the purchaser, and Mr. A. Lebrecht for the company, June, 1907. 
Motor cylinder 17X27^", normal speed 170 r.p.m. Fuel used was coke of the follow¬ 
ing composition: 


Moisture. 12.14 

Volatile matter. 3.05 

Fixed carbon. 77.58 

Ash. 7.23 

Heating value. 13 839 B.T.U. per lb. of dry coke. 


The engine was belted to an air compressor and run continuously for 72 hours at 
full load. Producer was charged regularly every hour, but only twice poked down and 
cleaned out during the run. The gas made carried considerable tar, and the tar 
extractors which were installed near the engine were cleaned at intervals of from 5-8 
hours without affecting the working of the engine. During such periods about \ lb. 







424 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


of tar would collect on the screens. The gas was analyzed from time to time, the 
result being shown in the following table: 


Sample. 

A 

B 

C 

D 

Date taken June, 1907—Time. 

10th, 

11.10 a.m. 

12th, 

3.20 p.m. 

13th, 

9.30 a.m. 

13th, 

2.30 p.m. 

Carbon dioxide, C0 2 . 

Oxygen, 0. 

Carbon monoxide, CO. 

Hydrogen, H 2 . 

Methane, CH 4 . 

Nitrogen, N. 

Heat value in B.T.U. per cubic foot. j 

6.20% 

.20 

26.80 

13.47 

0.10 

53.23 

142 

134 

5.60% 

3.00* 

23.60 

12.14 

0.10 

55.56 

126 

119 

6.20% 

.40 

27.80 

13.33 

none 

59.27 

143 

136 

6.00% 

1.20 

26.00 

14.40 

0.20 

52.20 

143 

135 


* Gas sample B happened to contain some air. 


On a preliminary run, during which' the engine was loaded down by a Prony 
brake, it was shown that for a B.H.P. of 75, the M.E.P. of the cylinder was 64.33 






lbs., equivalent to 85.7 I.H.P. This shows a mechanical efficiency of 87.5%. During 
the main test the load was obtained by an air compressor as above stated, the air 
being used around the works. Cards taken every 15 minutes during the test showed 
that the load was fully maintained, as may be seen by examining the cards of Fig. 
598 together with the table following. 






































STATIONARY ENGINES 


425 


Table 91 

CARD DATA FOR TEST OF 75 B.H.P. 4-CYCLE ENGINE 


Card. 

A 

B 

C 

D 

Date taken, June. 

10th 

11th 

12th 

13th 

Time, p.m. 

5.25 

4.45 

4.40 

12.45 

Revolutions per minute. 

172 

170 

171 

170 

M.E.P., lbs. 

64.3 

66.7 

66.0 

65.34 

I.H.P. 

88.46 

88.91 

88.50 

87.10 




The economic results were as follows: 


Duration of run, hours. 72 

B.H.P. developed. 75 

Total coke used, lbs. 5548 

Coke per hour, lbs. 77.06 

Coke appearing in ash, %. 6.7 

Coke per B.H.P. hour on test, lbs. 1.027 

Coke per B.H.P. hour, guaranteed, lbs. 1.200 

Water evaporated for producer, per B.H.P. hour, lbs. 1.13 


This plant possesses a relay unit of the same size. The engines are changed off 
every 24 hours, although in several instances the shift is not made until 48 hours 
have elapsed. The producers are changed off about once in three weeks for cleaning 
purposes. 

(6) Test of a 200 B.H.P. twin-cylinder gas engine at Petroleum Iron Works, 
Sharon, Pa., Dec. 28, 1907, by Mr. Kelly, chief engineer for the purchaser, and Mr. 
A. Lebrecht for the company. Motor cylinders 19lX3iy', normal speed 155 r.p.m. 
Direct connected to one 125 K.W. D.C. Western Electric generator. The engine will 
eventually run on producer gas, but natural gas is used as long as the supply lasts. 

This test is rendered somewhat unsatisfactory on account of the short time it 
lasted. The following table shows the principal results, the reading being taken every 
five minutes: 

Table 92 


Duration of test, 11:20 to 11:50 a.m., min.. 

Average load, K.W. 

Average load, E.H.P. 

Efficiency of generator, %. 

Average B.H.P. 

Average M.E.P., engine No. 78, lbs. 

Average M.E.P., engine No. 79, lbs. 

Average I.H.P., engine No. 78. 

Average I.H.P., engine No. 79. 

Total average I.H.P. 

Mechanical efficiency, %. 

Gas consumed per hr., engine No. 79, cu.ft. 
Gas consumed per hr., engine No. 78, cu.ft. 
Total gas consumption per hr., cu.ft. 


30 


rC0 2 . 

.20 

143.22 


O. 

1.70 

192.00 


CO. 

.10 

92.0 

Average composition . 

Illuminants. 

.40 

209.0 

of gas, vol. %, 

Hydrocarbon vapors 

3.40 

70.4 


CH,. 

91.80 

70.34 


H. 

.23 

125.8 


N. 

2.17 

128.0 

Heating value of gas, 

B.T.U. per f Higher 

1055.5 

253.8 

cu.ft., 

\ Lower 

951.5 

82.4 

Gas per B.H.P. hour (corrected for meter), 


Qoa 

r*n ft. . 


8.83 

oo O 

1012 

Consumption of heat per B.H.P. hr., B.T.U. 

8400 

1848 

Consumption of heat 

per B.H.P. per hr.. 



B.T.U., guaranteed. 

9000 


Consumption of heat per I.H.P. hr., B.T.U.. 

6920 


Thermal efficiency on I.H.P., %. 

36.8 


Thermal efficiency on B.H.P., % (economic 



efficiency). 


30.3 















































426 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


(c) Test of a 150 B.H.P. producer gas engine for Penn Hardware Co., Reading, 
Pa., Feb., 1907. 

Twin-cylinder engine 17X27£", normal speed 170 r.p.m., but on account of 
trouble with belt tested at 162 r.p.m. Belt connected to 60-cycle A. C. generator. 
The gas was furnished by a 300 H.P. “ Automatic ” pressure producer. The following 


table shows the main results: 

Table 93 

Duration of test, hours. 6 

Average B.H.P. 148.5 

Average I.H.P. 176.0 

Mechanical efficiency, %. 84.4 

Heating value of coal (higher), B.T.U. 11377 

Analysis of coal: Moisture, %. 1.12 

Volatile matter. 3.94 

Fixed carbon. 76.03 

Ash. 18.91 

Time of taking gas sample. 10 a.m. 2:45 p.m. 

Composition of gas, volume, %: C0 2 . 3.4 2.4 

O. 1.2 .50 

CO. 26.0 27.5 

H. 7.12 10.27 

CH 4 .70 .90 

N. 61.58 58.43 

Heat value of gas, B.T.U. per cu.ft.: Higher.116.6 135. 

Lower.113.1 128.5 

Coal consumed per B.H.P. hr., lbs. (including that in boiler). 1.24 

Gas per B.H.P. hr. (drop holder test), cu.ft. 73.8 

Heat consumed per B.H.P. hr. (in gas at 125.8 B.T.U.). 9500 

Heat consumed per B.H.P. hr. (in coal at 11377 B.T.U.). 14107 

Thermal efficiency of engine on I.H.P., %. 31.6 

Thermal efficiency of engine on B.H.P., %. 26.8 

Thermal efficiency of plant on B.H.P., %. 18.0 


The last two lines of the table show that the efficiency of the producer was 

18 

2 g-^ = 67.2%, which is fair considering that it was operated far below its normal load. 

The Hornsby-Akroyd Oil Engine. In these engines the compression pressure is only 
from 45 to 55 lbs., the maximum pressure from 180 to 200 lbs., the terminal pressure 
15 lbs., all by gauge. The M.E.P. at full load is from 40 to 42 lbs. 

(a) Tests given in catalogue of the De La Vergne Co. from “ Gas and Petroleum 
Engines,” by Robinson. 


Table 94 
TESTS OF 25 H.P. 


Load. 

Maximum. 

Full. 

Vs 

X 

No Load. 

Duration of trial, hours. 

x 

4 

3 

3 

2 

1 

R.p.m., mean. 

203 

202.6 

202.4 

203 

201.5 

M.E.P., pounds per square inch. 

. . . 

45.4, 43.4 

31.2 

18.3 

6 

I.H.P. 


32.3, 31.0 

22.4 

13.1 

4.28 

B.H.P. 

39 

26.74 

1 7 QR 

Q 


Mechanical efficiency. 

82.4, 86.0 

1 l . £7U 

80 

69 


Oil used per I.H.P., hour, lbs. 


0.61, 0.63 

0.74 

0.91 

1.34 

Oil used per B.H.P. hour, lbs. 


0.74 

0.91 

1.30 



(6) Test of a 20 B.H.P. engine made by Prof. W. Robinson, Feb., 1905. 

Normal speed of engine, 245 r.p.m., motor cylinder 11 xm". Fuel used was 
41 Russoline ” having a specific gravity of 0.825 at 60° F. 


















































STATIONARY ENGINES 


427 


Table 95 


RESULTS OF TRIALS OF HORNSBY OIL ENGINE (1905 TYPE) AT GRANTHAM, FEB. 23, 1905 


Power. 

Full. 

X 

X 

-- —, 

Light. 

Time taken to start. 

4 min. 
21.1 

26.4 

0.80 

13.5 

0.5 




Brake horse-power. 


10.75 

0 

Indicated horse-power. 

lo. yo 

Mechanical efficiency. 



6.3 

Oil Consumption. 

Total in engine, lbs. 

in ^ 

Q K 

5.25 

0.83 

Oil per indicated horse-power hour, lbs. 

IU . o 

O . O 

Oil per brake horse-power hour, lbs. 

0.64 

0.62 

0.66 

0.64 

A 7A 

Oil per brake horse-power hour, pints.. 

u. /y 
n 7A 

. 


u. /o 



No further information is given regarding the fuel oil, but assuming that its heating 
value is 18 500 B.T.U. per lb., the oil consumption at full load would correspond to an 


economic efficiency of =2L5 %* 

(c) Fig. 599 shows graphically the results of a fuel consumption test of an 85 
B.H.P. Hornsby-Akroyd engine made in Sept., 1906. 

The cylinder dimensions were 23X28", the normal speed 160 r.p.m. In this 
engine the piston was water-cooled. The fuel used was Standard Oil Co.’s fuel oil, 
but no further particulars with regard to fuel are given. The best consumption was 
apparently .84 lb. of oil per B.H.P. hour at 75 B.H.P. Assuming again that the oil 
has a heating value of 18 500 B.T.U. per pound, this figure would correspond to an 

2545 

economic efficiency of-= 16.4%. 

* .84X18500 /0 

The De La Vergne 2-cycle Vertical Oil Engine. The following averages at full 
load are representative of this type: compression 89 to 110 lbs., maximum pressure 
250 to 300 lbs., exhaust 50 lbs., M.E.P. 28 to 33 lbs., mechanical efficiency 75 to 80%. 

(a) Test of a 7X7£" twin cylinder engine, rated at 15 B.H.P. Engine direct 
connected to 10 K.W. Western Electric Co. generator on same base plate. Normal 
speed 500 r.p.m. Fuel used was kerosene. Tests made April 6 and 7, 1908. 

Table 96 


Load Factor. 

IX 

l 

l 

X 

X 

X 

©|o 

Length of test, minutes. 

15 

30 

30 

20 

20 

30 


500 

507 

520 

528 

535 

545 

Volts. 

126 

120 

120 

122 

120 

0 

Amperes. 

97 

83.5 

63 

44 

21 

0 

Kilowatts. 

12.22 

10.02 

7.56 

5.37 

2.52 

0 

Electrical H.P. 

16.37 

13.42 

10.13 

7.19 

3.38 

0 

Developed H.P. 

19.5 

15.8 

12.08 

8.88 

4.5 

0 

Generator efficiency, % . 

84 

85 

84 

81 

75 

0 

f 

Total oil used.^ 

3 lbs. 

5 lbs. 

4 lbs. 

2 lbs. 

lib. 

1 lb. 

15 oz. 

4 oz. 

9 oz. 

8 oz. 

13 oz. 

15 oz. 

Oil per hour, lbs. 

15.75 

10.5 

9.12 

7.5 

5.44 

3.88 

Oil per K.W. hour, lbs. 

1.3 

1.05 

1.2 

1.4 

2.16 

0 

Oil per E.H.P. hour, lbs. 

0.96 

0.78 

0.9 

1.04 

1.6 

0 

Oil per delivered H.P., lbs. 

0.81 

0.66 

0.76 

0.86 

1.21 

0 
























































428 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Fig. 599. 



















































































































































































STATIONARY ENGINES 


429 


The best consumption is 0.66 lbs. of oil per B.H.P. hour, which, with the assumption 
above made regarding the heating value of the oil, indicates an economic efficiency of 
2545 

•^g^g^Q = 20.8%. This figure, considering the high speed and small size of the engine, 
must be pronounced very good. 

3. The American Diesel Engine Co., New York City. 1 The Diesel engine, as 
made by the Vereinigte Maschinenfabrik Augsburg, represents the German type and the 
method of operation, together with some results obtained, have already been discussed 
on pp. 362 to 366, to which the reader is referred. This engine is to-day made by 
licensed manufacturers in nearly every country in Europe, and by the above named 
firm in this country. In all cases the engines built by the various manufacturers of 
course operate upon the principle already described, but the mechanical details of 
the various constructions differ here and there enough to make them distinct from 
each other. Thus the main point in which the American Diesel engine differs from 
the European types is in the use of the enclosed box frame. The others, to the 
writer’s knowledge, all use the open A-frame. 2 There are also the minor differences 
in the methods of obtaining compressed air, governor details, valve construction, etc. 

Plate XXXI shows the general assembly of a triple cylinder Diesel engine as 
built by the American Company. The base of the machine is a single casting com¬ 
pletely enclosed, so that splash lubrication may be employed for crank-shaft bearings, 
connecting rod bearings, cam shaft and cam rollers. The cylinders are cast in one 
piece with the jackets, the whole being secured to the box pedestal as shown. The 
cylinder head is a simple casting, containing a relief valve in the center of the head, 
while admission, exhaust, and fuel valves are located in a chamber at the side. Both 
admission and exhaust valves are of the simple poppet type. The admission valve 
works downward and the exhaust valve upward. The latter is not water-cooled. All 
■of the valves are operated by push rods, the exhaust valve directly, the admission 
valve through a lever pivoted to the admission valve housing and the fuel admission 
valve, which acts horizontally, by means of a bell-crank. All of this will be clear 
from a study of plate XXXI. 

Fig. 600 shows the valve construction in greater detail. The admission valve is 
held in a separate cage so that the seat may be easily taken care of. It closes 
against a small dash-pot located in the top of the housing for the purpose of reducing 
both noise and shock. The exhaust valve has no removable seat, but it is of such 
-dimensions that it may itself be easily removed through the admission valve cage 
opening, giving opportunity for grinding the seat whenever necessary. The relief valve 
in the center of the head is usually set to open at about 800 lbs. per square inch and 
acts as a safety against undue pressure caused by premature ignition, etc. The former 
may occur when the fuel valve has accidentally stuck, admitting oil to the cylinder 

during the compression stroke. . , 

The fuel valve, through which oil is admitted to the cylinder after the charge 
of air is compressed, is a very important part of the engine, and its construction is 
shown on a still larger scale in Fig. 601. Plate XXXI shows the location of the oil 
pump at the right hand side of the crank case and the manner of operating it by 


‘ Since the writing of this article, the American Diesel Engine Co. has been 

Bush. The main offices are now at St. Louis. 

’ See an article by R. Diesel, in the Zeitschnft d. V. D. Ing., 1903, p. 1368. 


bought by Mr. Adolphus 




430 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 

means of spur gearing from the cam shaft. From this pump an oil supply pipe leads 
to the oil valve shown at the left of section A-B, Fig. 601. From here the oil finds 
its way into the atomizer in the interior of the bushing held in the cast iron fuel 
valve cage. Both air and oil connections are screwed into this steel bushing so that 
the valve cage does not have to stand high pressure. The fuel admission valve itself 
consists of a nickel steel needle which carries a cast iron spring case on its outer end. 
Normally the spring forces this needle against its conical seat, but as the fuel-valve 
cam commences to operate, the bell crank shown pushes the needle to the right 
against the spring and opens the valve. This happens about the time that the 



main piston reaches the upper dead center on its compression stroke, and highly 
compressed air from a storage tank then rushes through the atomizer and forces the 
oil out into the compressed cylinder charge. In order to keep the needle cool and to 
keep the oil from carbonizing in the atomizer, water is circulated in the space between 
the steel bushing and the walls of the valve cage, both inlet and outlet pipes being 
shown in Fig. 601. 

Diesel engines are governed not by controlling the length of time that the fuel 
valve is open, but by adjusting the effective delivery stroke of the oil pump, and 
hence, except for minor variations in the setting of the fuel valve depending upon the 
kind of oil used, the lift and time of opening of all of the valves once set is always 
the same. Fig. 602 shows the method of setting the valves and the time of opening 
and closing, the lift being controlled by the fixed cams of the half-time shaft. 























































































PLATE XXXI 



ASSEMBLY □F 

TRIPLE CYLINDER ENGINE 

AMERICAN DIESEL ENGINE COMPANY N.Y 


stcT'on e-r 




tVSCOil** 


r //\/////////VA 


CARO 


'&ZZZZZ2. 














































































































































































































































































































































































































































































































































































































STATIONARY ENGINES 


431 

































































































































432 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



1 

£ 



© 

!S 

1 

!5 

co 


k 

o 

1 

s - 

to <? 

I 

Cj 

Q 

** 
k Vj 

{1 

1* 

to to £ 

|^<«* 

^a?s| 

«o 

VO 

NfVj 

°o 

inis 

2 to ^ "3 k! 

fO(^. 

<7> 

<Vj 

«Vj 

<*) 

> 

<r> 

^ Sfe «. ^ 

k o s 5 f 

sa^is 

<r> 

£ 

? 

Ni 

"5 

<o 

k 

O 

m 

«& t 

Sm| 

**S! K § 

VO 

£ 

> 

kB 

g 

'J- 

«D 

© 

k 

‘O 

s 

vo 

!s.SkS$ 
s?«i 25 
5 X k > 
^ v> to v> 

^"Vj 

> 

'o 

vo 

"«^ 

vo 

5 ? ^ to ^ 

1 ^ k! 5 s 

"IOj 

‘O 

*o 

NM 

N 


* 5 

V 1 


> 

CVj 

5 }- 


Is 

^ Vft 


<0 

<J) 


r '^ 

N 

k. !s 





X- 

o 5 


s! 

<\J 

<Vl 


kj 5? 


K 

* 

X 

K 

N ^ 


^l 

> 

VO 

VO 

to 




•v. 

N 


l>i s 

^ to * ^ 

£ *• <0 *) 

K V k k 

n f, > 5 

N t K ^ k. 

^ ^ ^ 

^ to to V kj 

N * N to * 

i ^8 ^ P 

8|fc*S! 

SlOto 

SS*** 

-> *t ? s 

to .to K < V> 

Sit*!? 

|SS;*s 


V. K 

*5 

. -g 


totok. 


s> to £ 


3'$ 

•> 2 V. 

5% 
i§S 
^ t $ 


. N» ^ XJ 

51 . ?.5s 


»s to NJ N K 

| ^ •> 

? ^ || $? to 

©| 'H 

2 * 5 to 
gc:$ $x§ 

% $ «o i to 


to to. ■ 
5 n to 
to to to 

v . to S 

Nl N ^ 

s * p 
s 5 

to K t 

**; 

*. v * 

^ ^' 
l big 
'’St 

S3* 

3^5 

*** 

SSI 


i ■* 2? 

$u 

^ ^ V| 

I °> 5 

£ •* k 
(o <* 

1 ^ 

^ ^ ^ 

K 2 

> ^ V* 

III 

ft K 5 

•• 5} *' 

» -> 

K v to 

k> c 
to 5 t 
•> > O 

, S to 

Sl’k 
3 * to ki' 
s* k ? 

3*52: 

51 ^ 

k < ^ 



Fig. 002. 





































































































STATIONARY ENGINES 


433 


The fuel pump with its governor is shown in Fig. 603. This shows a cross-section 
through one of the pump cylinders, there being as many pump cylinders as there are 
power cylinders. The pump plunger is actuated by a simple eccentric and strap, and 
its stroke is therefore constant. By means of a short horizontal arm and nearly 
vertical rod, the plunger is connected to a “pump suction valve eccentric lever” 
which, in conjunction with the fly-ball governor, controls the motion of the suction 
valve. At full load the governor sets the fulcrum about which the eccentric lever 
turns so that the suction valve opens when the plunger has completed half of its down 
stroke. The suction opening increases until the plunger has reached the lower end 
of its stroke. On the up stroke the valve is not closed until again half the stroke is 
completed, after which that part of the charge remaining is then forced through the 
double ball check valve and into the engine cylinder. The reason for keeping the 
valve open for half stroke each way is to let the oil free itself from air which it is 
very apt to entrain, and thus to deliver solid oil only. Under no load the governor 
so changes the motion of the eccentric lever that the suction valve is open practically 
during the entire up and down stroke of the plunger, so that little or no oil is 
delivered to the fuel injection valves. Between these two extremes of course the 
effective delivery stroke of the pump can be made anything to suit the load. 

The high pressure air used for injecting the fuel is usually obtained by inde¬ 
pendent compressors discharging into steel storage tanks. The air for starting is also 
obtained from this source. The pressure of the injection air should vary with the 
load on the engine, for half load or less it should be from 50-60 atmospheres, above 
that load from 65-70 atmospheres and for overloads 75 atmospheres is safe and 
permissible. 

The use of independent three-stage compressors in place of two-stage compressors 
driven from the engine is considered a distinct improvement over European practice. 
In a two-unit plant for instance, two independent compressors, one used as a relay, 
offer much greater security against a shut-down from lack of air than two engine 
driven compressors. This advantage is more pronounced as the number of units in 
the plant grows. Thus in a plant in Florida, three compressors serve to supply 12 
engine units, and the failure of any one of the three can not possibly effect the 
operation of the engines. Another important point is that such a system allows of a 
eool air supply to the injection valves, which helps to prevent the carbonizing of the 
fuel oil in this valve. 

To start the engine the fuel-valve cam on one of the cylinders is pulled over so that 
the fuel valve on that cylinder is closed and can not be opened. Instead of this the 
same operation brings into action a cam which controls a special starting valve on 
the same cylinder. The engine, after the crank-shaft has been brought into proper 
position, is then started by admitting compressed air to that cylinder. After one of 
the other cylinders is heard to obtain an ignition, the cam lever is returned to its 
former position, when the starting cylinder will also take up its regular cycle. 
Previous to starting the fuel pump must be operated by hand for a few turns by 
means of the starting wrench and pinion shown in Fig. 603, the pinion meshing with 
a gear on the pump shaft when the starting pin is pulled up (see the drawing of the 
cam shaft on Plate XXXI). The oil so pumped is discharged through an overflow, 
the purpose of this being to work all of the air out of the oil and to insure that nothing but 
solid oil is delivered to the fuel valves. 

The following table shows some details of Diesel engines at present regularly 
constructed by the American company: 


434 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 




Fig. 603. 






















































































































PLATE XXXII 

















































































































































































































































































































































































































































































































































































STATIONARY ENGINES 
Table 97 

ALL WEIGHTS ARE IN POUNDS 


435 


No. of 
Cylinders. 

Cylinder 

Size. 

R.p.m. 

Rated H.P. 

Weight 

without 

Fly-wheels. 

Number of 
Fly-wheels. 

Weight of 
each 

Fly-wheel. 

Total Weight 
of 

Fly-wheels. 

Total Weight 
of 

Engine. 

3 

3 

3 

3 

1 

1(H"X15" 
12" X18" 
14" X21" 
16" X24" 
16" X24" 

240 

220 

200 

164 

164 

75 

120 

170 

225 

75 

18500 

27000 

55000 

65000 

28600 

2 

2 

2 

1 

1 

2700 

5900 

6900 

15640 

11800 

5400 

11800 

13800 

15640 

11800 

23900 

38800 

68800 

80640 

40400 


A lay-out of two 225 H.P. three-cylinder engines driving a 300 K.W. generator, 
giving some further information regarding dimensions, is given in Fig. 604. In some 
installations of this type the rotor of the generator is made to act as a fly-wheel, 
which tends to shorten the length of the plant to a small extent. 

Finally, Plate XXXII gives a complete layout and piping plan for three 150 K.W. 
engines. 

Operating Results. (For results with Diesel engines of other makes, see p. 363.) 

The number of tests available on American Diesel engines is quite large, and 
consists partly of acceptance tests made usually by the company’s engineers, of 
operating reports received from a number of plants in commercial operation and a 
few tests made by independent observers. Except in one or two instances, nothing 
is stated concerning the heating value of the oil used, although the rest of the 
information is quite complete. Where efficiency computations are made, the writer 
has assumed the heating value of the oil at 19 000 B.T.U. per lb., except where 
otherwise stated. This is in most cases not far from truth, and in any case it is 
intended to show merely about what Diesel engines are doing as compared with 
other internal combustion engines. 

(a) Acceptance Test, Nov. 8-13, 1907, Perth Amboy. Triple-cylinder engine, 
16X24", 164 r.p.m., rated B.H.P., 225. Engine had two fly-wheels, one 8'6"X10", 
and one 7' 8"X8". Load obtained by brake on each wheel. Kind of oil, distillate. 


Load. 

Duration, Hours. 

R.p.m. 

B.H.P., Net. 

Oil per B.H.P. 
Hour, Lbs. 

Oil per 100 B.H.P. 
Hours, Gallons. 

Economic Effici¬ 
ency, Per Cent. 

i 

2 

173.5 

76.1 

.60 

8.4 

22.3 

I 

2 

168.6 

151.4 

.47 

6.6 

28.5 

Full 

2 

165.8 

227.7 

.45 

6.3 

29.8 

Full 

10 

164.8 

226.3 

.44 

6.2 

30.5 

Over 

4 

161.9 

257.2 

.43 

6.0 

31.3 


The engine did not drive its own air compressor. The allowances made at the 
different loads for this were as follows: 


i load. .. 
§ load. .. 
Full load 
Overload. 


17.6 H.P. 
18.4 
19.3 

19.7 



































436 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 




































































































































































STATIONARY ENGINES 


437 


(b) Acceptance test, Feb. 9-13, 1906, Beliefontaine. Triple-cylinder engine 

16X24", 164 r.p.m., rated B.H.P., 225. Load on engine obtained by generator belted 
to countershaft. Engine drove air compressor belted to countershaft. Kind of oil, 
distillate. 


Load. 

Duration. 

Hours. 

B.H.P. 

R.p.m. 

Oil, Pounds per 
B.H.P. Hour. 

Oil. Gallons per 
100 B.H.P. 
Hours. 

Efficiency of 
Generator 
and Shaft. 

Economic 

Efficiency, 

Per Cent. 


2 

90.5 

165.3 

.550 

7.58 

74.0 

24.2 

s 

2 

161.0 

164.0 

.463 

6.38 

83.5 

28.9 

Full 

2 

232.0 

163.0 

.453 

6.24 

86.5 

29.5 

Full 

10 

232.0 

162.0 

.445 

6.13 

86.5 

30.2 

Over 

4 

251.0 

161.7 

.441 

6.08 

87.0 

30.5 


Exhaust clear on all tests. 

(c) Acceptance test, May 2-5, 1906. Triple-cylinder engine, 16x24", 164 r.p.m. 
Kind of oil, distillate. 


Load. 

Duration, 

Hours. 

B.H.P. 

R.p.m. 

Oil, Pounds per 
B.H.P. Hour. 

Oil, Gallons per 
100 B.H.P. 
Hours. 

Economic 

Efficiency, 

Per Cent. 


2 

87.4 

173.9 

.558 

7.5 

23.9 

, § 

2 

156.7 

168.9 

.462 

6.2 

29.0 

Full 

2 

224.0 

164.8 

.452 

6.1 

29.7 

Full 

10 

224.0 

167.2 

.453 

6.1 

29.7 

Over 

4 

254.0 

163.6 

.459 

6.2 

29.2 


Engine drove air compressor. 

(ri) Acceptance test, Dec. 22-26, 1906. Triple engine, 16X24", 164 r.p.m. Kind 
of oil, Whiting fuel oil; weight, 7.42 lbs. per gallon. Heating value, 19 500 B.T.U. per pound. 


Load. 

Duration, 

Hours. 

B.H.P. 

R.p.m. 

Oil, Pounds per 
B.H.P. Hour. 

Oil, Gallons per 
100 B.H.P. 
Hours. 

Economic 

Efficiency, 

Per Cent. 

i 

2 

90.5 

171.2 

.556 

7.5 

23.5 

i 

2 

162.9 

166.8 

.481 

6.5 

27.3 

Full 

2 

230.3 

165.4 

.470 

6.33 

28.0 

Full 

8 

230.9 

165.9 

.477 

6.02 

29.5 

Over 

4 

258.5 

165.5 

.476 

6.42 

27.6 


Allowances, experimentally determined, made for air compressor which was driven 
by motor: 

^ load. 13 H.P. 

§ load. 15 

Full load. 22 

Overload. 2 5 








































438 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


Exhaust was clear on all tests. 

(e) Acceptance, July 30 and 31, 1907. Triple-cylinder engine, 16X24", 164 
r.p.m. Kind of oil, Whiting fuel oil, weight = 7.47 lbs. per gallon. Heating value, 
19 500 B.T.U. per pound. 


Load. 

Duration, 

Hours. 

B.H.P. 

R.p.m. 

Oil, Pounds per 
B.H.P. Hour. 

Oil, Gallons per 
100 B.H.P. 
Hours. 

Economic 

Efficiency, 

Per Cent. 


2 

83.1 

168.1 

.558 

7.47 

23.3 


2 

157.8 

166.8 

.443 

5.93 

28.7 

Full 

2 

228.4 

164.5 

.440 

5.89 

30.0 

Full 

8 

227.6 

164.1 

.452 

6.05 

29.2 

Over 

4 

254.9 

162.3 

.450 

6.02 

29.3 


Air compressor driven by motor. 

J load. 

§ load. 

Full load. 

Overload. 


Allowance made for this: 


17 H.P. 

19 

20 
21 


The average of the above five tests on the 16x24" three-cylinder engines seems to 
indicate that with crude oil we may expect the following results: 


Table 98 


Load. 

Oil, Pounds per B.H.P. 
Hour. 

Economic Efficiency, 

Per Cent. 

(Heat. Value taken at 
19000 B.T.U. per lb.) 


.564 

23.7 

i 

.464 

28.8 

Full 

.447 

30.1 

12% over 

.451 

29.7 


(/) Plant record, Traction Terminal Building, Indianapolis, Ind. Equipment: two 
170 B.H.P. engines, three-cylinder, 14X21", 200 r.p.m., two 120 K.W. D.C. generators. 


Total K.W. hours for January and February, 1906. . . . 105 230 

Total oil used for January and February. 11 009.5 gallons 

Average kilowatts per day. 1783 

Average oil per day. 185.4 gallons 

Average oil per 100 K.W. hours. 10.4 gallons 

Average oil per K.W. hour. . 104 gallons 

Average cost per K.W. hour, at 3 cents per gal. for fuel. 3£ mills 

Rated K.W. capacity of plant. 240 K.W. 

Average kilowatts per hour. 74.3 K.W. 

Average load factor. 30.54 % 


Load: 3 electric elevators; 15 arcs; 3000 inc. lamps connected; 100 H.P. in motors. 
( g ) Plant record, Citizens Electric Light & Ice Co., Lebanon, Ind. Equipment: 


One 225 B.H.P. 3-cyl. engine, 16X24", 165 r.p.m. 
One 120 B.H.P. 3-cyl. engine, 12X18", 225 r.p.m. 
One 150 K.W. A.C. generator 1 » , , . 

One 85 K.W. A.C. generator j 0perated ln parallel - 



























STATIONARY ENGINES 


439 


Plant operates 24 hours per day, but day load is very small, not exceeding 5 K W 
Results for period from Feb. 11, 1906, to Feb. 28, inclusive. 


Total K.W. hours. 27 850 

Total oil used. 3 339 gallons 

Oil per 100 K.W. 12 gallons 

Cost of fuel per K.W. 3^ mills 

Capacity of plant. 235 K.W. 

Average load. 63.6 K.W. 

Average load factor. 27% 

Load 

Number public arcs. 95 

Number private arcs. 29 

Number incandescent lamps connected. . . 6 000 

Motors on day circuit. 8 H.P. 


(h) Plant record, Wayne Paper Mills, Hartford City, Ind. Equipment: 

Two 225 B.H.P. 3-cyl. engines, 16X24", 165 r.p.m. 

Two 150 K.W. A.C. generators operated in parallel. 

Plant operates 144 hours per week. 

Record for period covering 18 consecutive days. 


Total K.W. hours. 91 753 

Total oil used. 8 831 gallons 

Average K.W. per day.. 5 609.6 

Average oil per day. 552 gallons 

Average oil per 100 K.W. 9.84 gallons 

Average oil per K.W. .0984 gallons 

Average cost per K.W. with oil at 3 cents per gal. ... 2.9 mills 

Rated capacity plant. 300 K.W. 

Average K.W. per hour. 233.33 

Average load factor.. 77.8% 


Cost of fuel per B.H.P. at engine, electrical efficiency 90% 2 mills 

Load consists of induction motors of various sizes. Total installation about 500 B.H.P. 
( i ) Acceptance tests made on a 170 B.H.P. three-cylinder engine at Camden, N. Y., 
after one month of service showed the following results: 1 


Load. 

Gallons of Oil per Hour 

Pounds of Oil per Hour 

Economic Efficiency, 
Per Cent. 

per K.W. 

per B.H.P. 

per K.W. 

per B.H.P. 

i 

.164 

.110 

1 21 

.815 

16.4 

i 

.112 

.076 

.83 

.562 

24.0 


.101 

.068 

.747 

.503 

26.9 

Full 

.101 

.068 

.747 

.503 

26.9 


1 Electrical World, June 2, 1906. 































440 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


To determine the brake horse-power from kilowatts, the writer of the article assumed 
a uniform efficiency of 90%. This is probably too low at the rated load and too 

high at the lower load. The last three columns have been added by the present 
writer assuming the oil to weigh 7.4 lbs. per gallon and that its heating value is 
19 000 B.T.U. per pound. 

(/c) Tests made on one 225 B.H.P. three-cylinder engine at the plant of the 

Kimberly & Clark paper mills, Kimberly, Wis. Engine drives one 150 K.W. A.C. 
generator. Tests were made by the Sargent Engineering Co. The figures in the 
following table give some idea of the fuel cost of compressing the air used for 

injection, together with some other data lacking in any of the other tests so far 

quoted. In this plant one unit of 120 B.H.P. was used to drive the air compressors 
for 900 B.H.P. of main engines. The table charges the 225 B.H.P. unit up with the 
proportionate part, that is one quarter, of the oil used in the compressor unit. 


Table 99 



X 

Load. 

X 

Load. 

M 

Load. 

Full. 

Load. 

Over- 

Load. 

Date of run. 

1-5-06 

1-5-06 

1-5-06 

1-13-06 

1-13-06 

Length of run in hours. 

1 


2 


2 


4 

.75 

Revolutions per minute. 

179 

.1 

177 

.1 

174 

.5 

169.1 

161 .2 

Brake horse-power. 

68 

. 15 

133 

.7 

194 

.9 

249.7 

263.8 

Electric horse-power. 

54 

.5 

113 

.6 

170 

.6 

224.8 

237.5 

Kilowatts. 

40 

.7 

84 

.8 

127 

.3 

167.7 

177.2 

Pounds of fuel oil used in main engine per hour. 

42 

.75 

57 

.1 

82 

5 

106.37 

126.66 

Pounds of fuel oil used in compressor unit per hour .... 

40 

.5 

35 

.6 

35 

.87 

38.93 

38.33 

One-quarte. oil used in compressor unit per hour. 

10 

.12 

8 

.9 

8 

.97 

.973 

9.58 

Total pounds of fuel oil used per hour. 

52 

.87 

66 

0 

91 

.47 

116.1 

136.24 

Pounds of fuel oil used per B.H.P. hour. 


.776 


.493 


.469 

.464 

.516- 

Pounds of fuel oil used per E.H.P. hour. . . . 


.97 


580 


53 

.516 

.573- 

Pounds of fuel oil used per K.W. hour. 

1 

29 


778 


718 

.69 

.77 

Pounds of jacket water per hour. 

2033 

5 

2528 

3 

3324 

3S94.5 

5306 

Pounds of jacket water per B.H.P. hour. 

29 

5 

18 

9 

17 

1 

15.6 

20.1 

Average temperature inlet water. 

34 

0 

34 

0 

34 

0 

38.8 

40.0 

Average temperature outlet water. 

126 

6 

135 

5 

137 

9 

180.0 

170.0 

B.T.U. per B.H.P. hour. 

15103 

9594 

9127 

9029 

10041 

B.T.U. per E.H.P. hour. 

18877 

11287 

10314 

10041 

11151 

2545 

Thermal eff. of engine and dynamo --—-- 

J B.T.U. per E.H.P." 

13 

4 

22 

5 

24 

6 

25.3 

22.8 

2545 

Thermal efficiency of engine - 

e B.T.U. per B.H.P. . 

16 

8 

26 

4 

27 

8 

28.1 

25.3 

Percentage B.T.U. in water jacket. 

18 

9 

20 

0 

19 

5 

24.3 

26 0 

Percentage B.T.U. in exhaust, radiation and engine 







friction. 

65 

0 

53 

4 

52 

7 

47.6 

48 7 

Percentage B.T.U. in B.H.P 

16 

8 

26 

4 

27 

8 

28.1 

25 3 

Pounds of oil per 100 B.H.P. hours . 

77 

6 

49 

3 

46 

9 

46.4 

51 6 

Gallons oil per 100 B.H.P. hours 

10 

6 

6 

7 

6 

4 

6 3 

7 0 

Gallons oil per 100 K.W. hours . 

17 

6 

10 

6 

9 

8 

9 4 

10.5 

Fuel cost of 100 B.H.P. hours in cents with oil at 3.64 







cents per gallon. 

38 

6 

24 

5 

23 

3 

23 0 

25 6 

Fuel cost of 100 K.W. hours, cents 

63. 

0 

38 

5 

35 

6 

34.2 

38.2 


(l) 1 Tests made on the plant of the Prairie Pebble Phosphate Co., Mulberry, Fla., 
by Mr. R. E. Ludwig, between Feb. 7 and July 2, 1908. 


‘See Electrical World, Oct. 3, 1908, for description of plant. 























































STATIONARY ENGINES 


441 


This plant consists of 7 units of 450 B.H.P., each direct-connected to a 300 K.W. 
Allis-Chalmers-Bullock alternating current generator. Each unit is composed of two 225 
B.H.P. engines, standard dimensions, as shown on Plate XXXI. The generator is of 
the three-phase revolving field type operating at 60 cycles, 2300 volts, normal speed is 
164 r.p.m. Each engine also drives a belted exciter. 

The compressed air is furnished by five compressor units of the three-stage type, 
each unit being geared to a 75 H.P. Allis-Chalmers induction motor. 

Each unit was tested separately, the load being obtained by water rheostat. The 
compressors were not driven by the unit under test but from the general bus-bars. 

The oil used was desulphurized Texas fuel oil having the following characteristics. 


Sp.gr. at 60° F. .8689 

Wt. per gallon at 60° F., lbs. 7.238 

Flash point, ° F. 150 

Water. None 

Sulphur, %. .17 

Acidity. None 

Heating value, B.T.U. per lb. 19521 


The oil consumption tests on Units 2, 4, 6 and 7, resulted as follows: 


• Unit No. 

Oil Used ; 

per Hour. 

Oil Used per 100 K.W. Hours. 

Full Load (315 K.W.). 

Half Load (160 K.W.). 

Full Load (315 K.W.). 

Half Load (160 K.W.). 

2 

196.6 lbs. 

100.25 lbs. 

8.62 gals. 

8.66 gals. 

4 

187.38 “ 

102.8 “ 

8.22 “ 

8.88 “ 

6 

195.5 “ 

102.0 “ 

8.57 “ 

8.81 “ 

7 

187.0 “ 

102.5 " 

8.20 “ 

8.85 " 

Average 

191.63 lbs. 

101.89 lbs. 

8.403 gals. 

8.80 gals. 


The efficiency of the generator was given equal to 93.3% at full load, and 90.10% 
at half load. The quantities for this plant were based on brake horse-power and 
cover the oil consumption of the compressor, but not that of the exciter. Hence the 
figures in the above table must be corrected. 

The compressor power was found to be equal to 30 H.P. The exciter power 
consumption was determined at 8.7 K.W. or 13.7 H.P. The engine unit must be 
credited with the latter amount but should be charged up with the former. The net 
deduction is therefore 30-13.7 = 16.3 H.P. = 12.2 K.W. at the switchboard. The same 
allowance was made for both full and half loads. With the aid of the above generator 
efficiencies the results were next reduced to the contract basis with the following 
showing: 


Load. 

Net B.H.P. 

Gallons of Oil per 100 B.H.P. Hours 

Pounds of Oil 

P U p TTrm r 

Economic 

Efficiency. 

From Tests. 

Guaranteed. 

per D.n.r. nour 

from Tests. 

Per Cent. 

315 K.W., Full. 

435.4 

6.08 

8.0 

.440 

29.7 

160 K.W., Half. 

219.9 

6.40 

9.5 

.465 

28.0 







































442 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


The last two columns were added by the writer by aid of the data given above 
for the oil used. 

Unit No. 4 was also tested at various other loads with the following results. 


Load, K.W. 

Oil Used per 
Hour, Lbs. 

Oil per 100 K.W. 
Hours, Gallons. 

Net B.H.P.’ 

Oil per 100 
B.H.P. Hours, 
Gallons. 

Oil per B.H.P. 
Hour, Lbs. 

Economic 

Efficiency, 

Per Cent. 

0 

80 

56.00 

77.33 

13.36 

107.4 

9.94 

.720 

18.2 

160 

102.80 

8.88 

219.9 

6.46 

.468 

27.8 

240 

147.33 

8.48 

330.8 

6.16 

.447 

29.2 

315 

187.38 

8.22 

435.4 

5.95 

.432 

30.2 

330 

202.00 

8.46 

456.6 

6.11 

.443 

29.6 

340 

211.00 

8.58 

470.0 

6.20 

.449 

29.0 


Tests were also made to determine capacity and regulation. The average of all 
the capacity tests on the 7 units showed that the average maximum load carried 
was 473 B.H.P., which corresponds to an overload capacity of 5%. The regulation 
tests apparently consisted merely of determining the full and no-load speeds, the total 
difference between these two extremes was guaranteed not to exceed 10 r.p.m., or 
about 6%. The average total difference found on the tests was 8.5 r.p.m. or about 
5.2%, all of the engines individually meeting the guarantee. 

(m) Tests made on a 450 B.H.P. unit direct-connected to a 330 K.W. A.C. 
generator in the Silver Lake Power Plant at Pittsfield, Mass. The tests were made 
under the direction of Sibley College, Cornell University, by Messrs. A. Kennedy, 
J. L. Robbins, and M. duP. Lee, in April, 1907. 

The auxiliary machinery consisted of a 15 K.W. D.C. exciter belted to an A.C. 
Stanley electric motor of 33 H.P.; two three-stage Norwalk air compressors belt 

driven by 25 H.P. Stanley electric motors; and a triplex Gould pump belted to a 
5 H.P. A.C. motor, to supply jacket water. 

The engines were of the same general dimensions as the 225 B.H.P. machines 

already described. For the purpose of the test, one half of the unit was disconnected 

and the other loaded by the generator, putting on service circuits as required. This 
of course made the generator operate at all times far below full load, but for the 
purpose of this test the generator was used merely as a brake, a calibration curve 
furnished by the Stanley Co. being used to reduce the K.W. output to the horse¬ 
power basis. 

The crude oil used on the test was examined by Mr. B. S. Cushman of the 

Chemistry Department of the University, and showed the following characteristics: 


Composition, C =86.97% by weight 
H = 12.50% 

S= .30% 

0= .23% 

Heating value = 19 272 B.T.U. per lb. 

Sp. Gravity at 60° F. = . 8776 

The following table shows the results of the tests in detail, and but little explana¬ 
tion is needed. The indicators gave unreliable results when the engine was running 
at about £ load, and this data is therefore not recorded. The reason for this appar- 














STATIONARY ENGINES 


443 


ently was that each cylinder obtained an ignition only now and then, the injection 
air driving the cylinder the rest of the time. As a result the indicator cards at this 
load showed scattered expansion lines, and it was not possible to get the I.H.P. 
accurately. For the remainder of the tests the cards were quite normal. 

During normal operation the water used to cool the oil-injection valves was sprayed into 
the exhaust pipe for the purpose of keeping it cool. In order to determine the real 
exhaust gas temperature, this water supply was cut off several times for as long a period 
as appeared safe, hence the two lines recording exhaust gas temperature in the table. 

An attempt was made to determine in two ways the ratio of air to oil used, 
from the analysis of the exhaust gas in conjunction with the analysis of the oil 
and by means of the volumetric efficiency of the cylinder as found from lower loop cards 
taken on each run. The results of these two methods of computation are also given 
in the table. Neither method is quite correct, since on the one hand uncertainty 
exists as to the temperature in the cylinder at the end of the suction stroke, while 
on the other no means was at hand to analyze the exhaust gases for unburned fuel. 
Attention is called, however, to the fact that at the high loads the agreement is close. 
The air theoretically required to burn each pound of the kind of oil used was 15.3 
lbs., hence the air excess in operation was at least 55% at the highest load. 

Table 100 



I 

II 

III 

IV 

V 

VI 

Duration, minutes. 

100 

90 

90 

70 

110 

60 

Barometer, inches Hg. 

29.28 

29.28 

29.28 

29.28 

29.28 

29.28 

Output generator, K.W. 

26.61 

58.53 

85.39 

116.26 

146.37 

162.51 

Volts. 

2362 

2424 

2418 

2443 

2450 

2438 

Amperes. 

20.16 

33.35 

45.51 

63.72 

76.49 

81.46 

Power factor, %. 

55.7 

69.0 

77.8 

74.5 

78.2 

81 .8 

Generator friction, wind., etc., K.W.. . 

9.8 

10.2 

10.1 

10.4 

10.42 

10.4 

Armature loss, K.W. 

.104 

.29 

.544 

1.08 

1.56 

1.77 

Total generator loss, K.W. 

9.9 

10.49 

10.64 

11.48 

11.98 

12.17 

Efficiency generator, %. 

72.8 

84.3 

88.5 

90 5 

91 .6 

92.0 

Engine output, gross, K.W. 

36.51 

69.02 

96.03 

127.74 

158.35 

174.68 

Engine output, gross, H.P. 

49.0 

92.5 

128.9 

171 .0 

212.2 

234.0 

Imput. comp, motor, K.W. 

16.0 

15.58 

15.54 

15.6 

15.8 

15.54 

Efficiency comp, motor, %. 

87.5 

87.5 

87.5 

87.5 

87.5 

87.5 

Output comp, motor, K.W. 

14.0 

13.62 

13.6 

13 64 

13.8 

13.6 

Input pump motor, K.W. 

2.79 

2.72 

2.8 

2.9 

2.8 

2.7 

Efficiency pump motor, %. 

80 

80 

80 

80 

80 

80 

Output pump motor, K.W. 

2.23 

2.18 

2.24 

2.32 

2.24 

2.16 

Out. comp, and pump motor, K.W. . . 

16.23 

15.8 

15.84 

15.96 

16.04 

15.76 

Engine output, net K.W. 

20.27 

53.22 

80.19 

111.78 

142.31 

158.92 

Engine output, net H.P. 

27.15 

71.4 

109.7 

149.6 

191.0 

213.0 

Jacket water, total cu.ft. 

68.4 

89.6 

113.1 

96.8 

201.0 

69.0 

Jacket water, total, lbs. 

4260 

. 5600 

7060 

6040 

12550 

4300 

Pounds water per hour. 

2560 

3735 

4710 

5175 

6850 

4300 

Temperature water entering, ° F. 

50.3 

51 .53 

52 

52 

52 

53 .47 

Temperature water leaving. 

108 

110.7 

114.45 

119.3 

124.1 

179.6 

Water temperature range. 

57.7 

59.17 

62.45 

67.3 

72.1 

126.13 

Temperature room and supply air .... 

66.5 

67.9 

68.15 

68.15 

68.2 

69.3 

Temperature exhaust gas with water . 

255.0 

345.0 

375.0 

465 0 

498 0 

645 .0 

Exhaust gas range with water, ° F . . . 

188.5 

277.1 

306.85 

396.85 

429.8 

575 7 

Temperature exhaust gas, no water. . . 

257.0 

360.0 

412.0 

525.0 

642.0 

705 0 

Exhaust gas, range, no water, ° F. 

190.5 

292.1 

343.85 

456.85 

573.8 

635. i 

Specific heat exhaust gas. 

.2410 

.2418 

.2432 

.2442 

.2457 

. 2479 

Oil, total lbs. 

63.06 

67.1 

85.75 

86.1 

166.0 

108.3 

Oil per minute, lbs. 

.6306 

.7455 

.95277 

1.23 

1.509 

1 .805 

Oil per hour, lbs. 

37.84 

44.733 

57.166 

73.8 

90.54 

108.3 

Revolutions per minute. 

167.57 

165.8 

166.0 

167.0 

167.7 

167i0 

Air tank pressure in atmospheres. 

65.9 

67.2 

68.0 

71 8 

73.5 

71.85 

Air tank pressure, lbs. 

969 

989 

1000 

1055 

1080 

1057 

















































444 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 

Table 100 —Continued 



air 

Ratio — by weight, from exhaust gas 
oil 

analysis. 

Same, from vol. eff. 


84 26 
69.92 


Indicator number. 

Maximum pressure, lbs. persq. in 


1 2 
6401494 


3 

568 


M.E.P. 

I.H.P. per cylinder 


Total I.H.P. 
D.H.P. gross 
D.H.P. net . 


49.0 

27.15 


Mechanical efficiency, gross. 

Mechanical efficiency, net. 

Friction H.P. 

Oil per I.H.P. per hour, gals. 

Oil per I.H.P. per hour, lbs. 

Oil per gross D.H.P. per hour, gals. . . 

Oil per net D.H.P. per hour, gals. 

Oil per gross D.H.P. per hour, lbs.*. . . 

Oil per net D.H.P. per hour, lbs. 

Oil per 100 net D.H.P. per hour, gals. . 
Fuel cost per 100 D.H.P. hour, cents. . 
Heat balance: 

Heat supplied per hour, B.T.U. 

Heat supplied per hour, % . 

Absorbed in jacket, B.T.U. 

Absorbed in jacket, %. 

Exhaust, B.T.U. 

Exhaust, %. 

Thermal equivalent indicated work . 
Thermal equiv. indicated work, % ■ ■ 

Radiation and loss, B.T.U. 

Radiation and loss, %....'. 

B.T.U. per I.H.P. per hour. 

B.T.U. per gross D.H.P. per hour .. . 
B.T.U. per net D.H.P. per hour .... 

Thermal efficiency, I.H.P. 

Thermal efficiency, gross D.H.P. . . 

Thermal efficiency, net D.H.P. 

Volumetric efficiency, % . 


.1055 
.19 
. 77 
1.385 
19.0 
68.5 

729000 

100 

147800 

20.3 

123100 

16.9 


14900 

26700 


17.2 

9.55 

95.5 


II 

III 

IV 

V 

IV 


55.67 


50.30 


41.44 


26.44 

24.00 


55.76 


42.86 


34.61 


28.88 

23.93 

1 

2 

3 

1 

2 

3 

1 

2 

3 

1 

2 I 3 

1 

2 

3 

620 

540 

572 

630 

521 

573 

640 

536 

601 

640 

53 ^ 1 606 

704 

590 

640 

38.5 

30.0 

40.28 

57.5 

37.77 

44.3 

72.3 

55 0 

51.9 

83.7 

66.3 68.9 

90.2 

79.8 

73 

38.9 

27.5 

40.7 

52.8 

38.16 

44.7 

73.5 

55.9 

52.7 

85.2 

67.e|70.2 

92.1 

81 .4 

74 

107.09 

141.1 


182.15 


223.0 

248.30 


92.5 


128.9 


17 

1.0 


212.0 

234.0 



71.4 


109.7 


149.6 


191.0 

213.0 


86.5 


91.3 


94.0 


95.1 

94.4 


66.2 



77.6 


82.5 


85.6 

86.0 



14.59 


12.2 


11.15 


10.8 

14.3 



.0571 


.0555 


.0555 


.0557 


.0597 


.417 


.405 


.405 


.406 


.437 



.066 


.0604 


.0586 


.0578 


.0627 


.0853 


.0723 


.0669 


.6045 


.069 


.482 


.441 


.427 


.4225 


.458 


.623 


53 



.488 


.470 


.504 



8.53 


7.23 



6.69 



6.45 

6.9 


30.7 


26.0 


24.05 


23.2 

24.8 


864000 

1100000 

1420000 

1745000 

2080000 


100 



100 



100 



100 


100 


220000 


294000 

347000 

293000 

544000 



25., 



26.7 



24.4 



28.2 

26.2 



182000 


208000 

291000 

378000 

421000 



21. 

L 


18.9 



20.5 



21.7 

20.2 


273000 


359000 

463000 

567000 

632000 



31. 



32.6 



32.6 



32.5 

30.4 


189000 


239000 

319000 

306500 

483000 



21.8 


21.8 



22.5 



17.6 

23.2 



8060 


7800 



7800 



7820 

8380 



9310 


8530 



8240 



8150 

» 8820 



12000 


10180 



9400 



9100 

9680 



31. 



32.6 



32.6 



32.5 

30.4 



27.35 


30.0 



30.9 



31.2 

28.8 



21.20 


25.0 



27.0 



28.1 

26.2 



92. 

5 


89.0 



92.6 



94.4 

93 6 



* Oil at 8.3 cents per gallon. 


(n) Acceptance tests, March, 1908, Waco, Texas. 

Three-cylinder 14X21" engine rated at 170 B.H.P., r.p.m., 200. Engine had two 
wheels, 7 ft. X 18X14", each carrying a brake for the tests. Air compressor independently 
driven. Allowance made for this, 16 H.P. Kind of oil used, distillate, sp.gr. =31° B. 


Load. 

Duration, 

Hours. 

R.p.m. 

B.H.P., Net. 

Gallons of Oil 
per 100 B.H.P. 
Hours. 

Pounds of Oil 
per B.H.P. 
Hour. 

Economic 

Efficiency, 

Per Cent. 

* 

2 

209.3 

58.3 

9.3 

.68 


§ 

2 

205.5 

117.G 

7.3 

.53 


Full 

2 

202.5 

172.3 

6.4 

.46 


Full 

10 

202.1 

172.1 

G .4 

.47 


Over 

4 

199.3 

190.3 

6.6 

.48 














































































STATIONARY ENGINES 


445 


(o) Acceptance tests, August, 1907, Coeymans, N. Y. 

Three-cylinder 12X18" engine rated at 120 B.H.P., 220 r.p.m., Engine had two 
84X12" wheels, one carrying brake for tests. Air compressor independently operated. 
Allowances made were: 

For 4 and § load 

For full load. 

For overload. . . . 


Kind of oil, distillate. 


Load. 

Duration, 

Hours. 

R.p.m. 

B.H.P., Net. 

i 

Gallons of Oil 
per 100 B.H.P. 
Hours. 

Pounds of Oil 
per B.H.P. 
Hour. 

Economic 

Efficiency, 

Per Cent. 

4 

2 

228.8 

39.3 

9.6 

.69 


4 

2 

226.6 

82.0 

6.9 

.49 


Full 

2 

220.6 

120.9 

6.7 

.48 


Full 

10 

222.2 

121.9 

6.6 

.47 


Over 

4 

218.9 

135.4 

6.4 

.46 



16 H.P. 
16.3 H.P. 
16.5 H.P. 


4. The Snow Steam Pump Works, Buffalo, N. Y. This company, the first 
in the United States to undertake the building of large gas engines, confines its 
product exclusively to the double-acting 4-cycle horizontal engine. Two types of 
these are built, Type B up to 500 B.H.P. in single tandem and up to 1000 B.H.P. 
in twin-tandem units, and type A up to 2500 B.H.P. in single-tandem and up to 
5000 B.H.P. in twin-tandem units. Type B engine is also furnished as a single¬ 
cylinder or twin-cylinder engine. The following table shows the principal dimensions of the 
two types. Any of these sizes are built to operate on any of the following fuels: natural, 
producer, blast furnace, illuminating, coke oven, and all other industrial power gases. 


Table 101 

PRINCIPAL DIMENSIONS OF SNOW ENGINES 


i 

No. 

Type. 

B.H.P. 

Diameter Cylinder. 

Stroke. 

R.p.m. 

S.T. 

T.T. 

Natural Gas. 

Producer Gas. 

4 

B 

60 

120 

11" 

124" 

12" 

250 

5 

B 

80 

160 

1H 

124 

15 

225 

6 

B 

100 

200 

124 

14 

15 

225 

7 

B 

125 

250 

134 

14! 

18 

208 

8 

B 

150 

300 

14 

154 

21 

197 

9 

B 

200 

400 

154 

17 

24 

185 

10 

B 

250 

500 

16 

18 

30 

150 

11 

B 

375 

750 

194 

22 

36 

130 

12 

B 

500 

1000 

22 

24 

36 

130 

8 

A 

400 

800 

22 

24 

36 

110 

84 

A 

500 

1000 

23 

254 

42 

105 

9 

A 

650 

1300 

254 

28 

42 

105 

10 

A 

800 

1600 

28 

31 

42 

105 

11 

A 

1000 

2000 

30 

32 

48 

100 

12 

A 

1250 

2500 

33 

37 

48 

100 

13 

A 

1500 

3000 

35 

39 

54 

95 

14 

A 

1750 

3500 

37 

41 

54 

95 

15 

A 

2000 

4000 

40 

44 

60 

90 

154 

A 

2250 

4500 

42 

47 

60 

90 

16 

A 

2500 

5000 

43 

48 

60 

90 


* 





































446 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


There are two features in which the design of both types of engines differs radically 
from conventional European practice. The first is the adoption of the side crank in place 
of the center crank frame, and the second that the valves all open into a chamber at 
the side of the cylinder. All inlet valves and their gear are placed on top of these 
chambers and all exhaust valves and gearing on the bottom. The great advantage of 
this construction is that no part of the valve gear is below the floor, and the center 
of the cylinders can be kept- low, making the engine rigid and steady. The foundation 
can be one continuous block instead of a series of isolated pieces. The foundation 
plate is continuous under all of the cylinders, and the latter slide upon this plate on 



Fig. 605. 


machined surfaces, thus allowing of free expansion and contraction. The method of 
locating the valves in this manner makes the valve gear very simple, as may be 
seen from any of the following illustrations showing the gear. But one cam is used 
to operate both inlet and exhaust valves. The fear expressed that the location of 
the valve chamber at the side would effect the efficiency of combustion in an unfavor¬ 
able sense, has apparently not materialized in practice. The further objection made 
that the cylinder as constructed would not be able to clear itself of any foreign material, 
such as dirt from the gas, incrusted oil, etc., also does not appear to have much 
weight. 

The constructive features of one of the small sized Type B engines are given in 
Figs. 605 and 606. For the larger sizes of this type the valve-gear construction is 
somewhat different, as may be seen from a study of Plates XXXIII, XXXIV, XXXV 
and Fig. 607. It will be noted that the speed regulation of this engine is effected by 






PLATE XXXIII 



Snow Engine, Type B 






















































































































































































































































* 












































Snow Engine, Type B 
















































































































PLATE XXXV 



Snow Engine, Type B 




































PLATE XXXVI 



Type A Snow Engine, 400-5000 H.P. 
















































































































































• 











































' 
























































• 

















































■ 

















































































I B :' ■ 

■ 


























STATIONARY ENGINES 


447 


Fig. 606. —Valve Gear, Smaller Sizes of Type B Snow Engines. 




Fig. 607.—Valve Gear, Type B Snow Engines. 200 to 1000 H.P. 





























448 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


a centrifugal governor which controls a combined mixing and throttling valve. The 
latter is placed midway between the two inlet valves and serves both of them. 

Type A engines differ from Type B in many points of frame and cylinder design, 
as will be seen by comparing Plate XXXVI with Plate XXXV. The general features 
of the valve gear on the two engines are the same except that the construction of the 
regulating and mixing valve in Type A differs with the service to which the engine is 
to be put. The first large installation of Snow engines consisted of four 5400 H.P. 
gas engines for the California Gas and Electric Corporation, San Francisco. One of 
these units which were used to operate Crocker-Wheeler generators, is shown in Fig. 
608. In these engines the mixing and regulating valve was a sleeve valve concentric 
with the main inlet valve. Since then the use of this valve has been abandoned in 
favor of other constructions. Where the engine is to be used for a service which 
does not call for great refinement in speed regulation, such as pumping gas and 
water, or compressing air, the inlet' valve gear of Type A is the same as that of Type 
B. Where close regulation is required, however as in electrical service, a separate drop 
cut-off mixing and regulating valve is placed just ahead of each inlet valve. The 
former construction of Type A gear (also Type B) is very well shown in Fig. 609, 
which shows three 1100 H.P. gas compressors, while the operation of the drop cut-off 
gear may be explained by the use of Figs. 610 and 611. 1 In the former figure, the 
air is admitted through A and the gas through G to the mixing chamber M above 
the main inlet valve I. Disks A and G, together with the short barrel D, form one 
casting. The disk G is provided with a taper seat and the length of the valve stem 
is adjusted in the block at the upper end so that both A and G will seat at the same 
time. The ratio of air to gas in the mixture is set once for all by adjusting the 
baffle diskR by means of the knurled nut N. The cut-off valve is operated as follows: 
Inlet valve I is opened by the rocker arm, Fig. 611, always at the beginning and end 
of the suction stroke. By means of the link R, Fig. 610, the main inlet valve 
stem operates a sliding block in a guide above the cut-off valve. To this sliding 
block is pivoted a latch, L, Fig. 611, which, when in position, engages with the 
block on the end of the cut-off valve stem and thus- lifts this valve when the inlet 
valve I opens. To close the cut-off valve the governor controls the latch L, dragging 
it out of position by means of the drag link and the cam C shown in Fig. 611. The 
right hand end of the drag link curves around the governor shaft S and rests upon 
the journal box, thus holding the link in place as it slides back and forth. The cam 

C engages a lug on this link, thus displacing it and the latch L. At the moment the 

latter frees the block, the cut-off valve is closed by means of its spring, Fig. 610, and 
no more mixture is admitted to the cylinder. The cam shaft S rotates continuously 
at half speed. The point in the suction stroke at which cut-off occurs depends entirely 
upon the position of the cam, C, with relation to the crank-shaft, and this is con¬ 
trolled by the governor through a so-called “ floating ” bevel gear. 

Operating Results. Figures relating to the performance of Snow engines in actual 
service are extremely scarce. The only results that the writer has ever seen were 

given in Power, July 14, 1908, and these do not give all of the data desirable. 

Double-acting twin tandem engine at Ceres, N. Y. Cylinders 16" diameter X 30" 
stroke, 150 r.p.m., driving 300 K.W. alternator. Fuel is natural gas estimated at 900 
B.T.U. per cu.ft. The following table shows the gas and heat consumption for one 


1 Power, July 14, 1908. 



STATIONARY ENGINES 
Fig. 609. 


449 



Fig. 608 





























450 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Fig. 610. 



Fig. 611 . 













































































































STATIONARY ENGINES 


451 


week. In view of the fact that the average load was only 240.8 K.W. the average 
heat consumption of 14 186 B.T.U. per B.H.P. hour, which corresponds to an economic 
efficiency of 17.9%, is not bad. To the table from Power , the writer has added the 
last two columns. 


Table 102 


Day. 

1 

Hours 

Run. 

Kilowatts. 

Average 

Load, 

Kilowatts. 

Total 

Cubic Feet 
of Gas. 

Gas per 
Kilowatt 
Hour, 
Cubic Feet. 

B.T.U. 
per Brake 
H.P. 
Hour.* 

Economic 
Efficiency, 
Per Cent. 

Load, 

Per Cent 
of Full. 

Sunday . 

21 

4040 

192.4 

109960 

27.2 

16320 

15.5 

64.0 

Monday. 

20 

5050 

202.5 

127530 

25.2 

15120 

16.8 

67.5 

Tuesday. 

21 

4890 

232.9 

116400 

23.8 

14280 

17.8 

77.4 

Wednesday. 

19 

4730 

248.9 

106310 

22.5 

13500 

18.9 

83.0 

Thursday. 

18 

5050 

280.5 

111610 

22.1 

13260 

19.2 

93.2 

Friday. 

20 

5120 

256.0 

114060 

22.5 

13500 

18.9 

85.0 

Saturday. 

19 

5180 

272.6 

115060 

22.2 

13320 

19.1 

91.0 


♦Assuming 900 B.T.U per cubic foot and brake horse-power per kilowatt. 


5. The Olds Gas Power Co., Lansing, Mich. Type “ K ” Engines for Pro¬ 
ducer, Natural and Blast Furnace Gas. In the production of an engine for the heavier 
gas power service, the Olds Company have adhered to the single-cylinder, single-acting, 
horizontal type of 4-cycle machine for all units up to 150 H.P. Also above this 
limit, a large number of plants have been installed, using a set of two or more of 
these machines, the reason for this policy being that in constructing engines of this, 
type, absolutely no uncertain elements are introduced into the design. The various 
details of construction of this line have been repeatedly justified by their own 
engines in practical operation, and also by thousands of similar machines abroad. 

As is usual with this type of construction, the cylinder jacket, or casting, is cast 
integral with the engine frame. From each side of the cylinder jacket casting two 
wing-like box girders extend to the main bearings. The outlines of this frame are 
well shown in the general view, Fig. 612 and the sectional view, Fig. 613. 

Again, proceeding along well-tried lines, the cylinder proper is a separate casting, 
which is forced into the main casting from the head end. The advantages of thus 
being able to use a specially hard, close-grained iron for this particular casting, and 
also giving the liner perfect freedom to expand at the end, are now well understood 
in this country. 

The cylinder head is still another casting containing the main valves, but for the 
sake of simplicity and strength contains as little of their various passages as practical. 

The Olds Company anneal all of these heads to insure that initial shrinkage 
stresses are removed. As indicated by the above views, the construction is amply 
massive to withstand the heavy strains of the high-compression cycle. 

A desirable feature which can always be attained in this construction is the 
absence of overhang of the cylinder. Thus, that part of the cylinder which takes the 
side thrust of the connecting rod is well within the bed. 

In the larger sizes not only are the cylinder and head jacketed, but the exhaust 
valve and pipe as well. 

There are a number of excellent design features embodied in these machines, 
among which may be mentioned the following. 
























452 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 

The crank-shaft is an open-hearth steel forging, with the crank slotted from the 
solid. In the design of the shaft, it has been the policy of this firm to avoid, as 
far as possible, the overhung fly-wheel, and accordingly their standard machine has 
but one fly-wheel, which is designed of ample proportions to give the necessary cyclic 
regularity; and the shaft is supported by three bearings—two main bearings in the 
engine bed, and one bearing outside of the fly-wheel. The counter-balances for the 
cranks are both keyed and bolted to the cheeks. The shaft is of enough greater 
diameter at the fly-wheel to admit of key-seating without reducing the strength of 
the shaft. Two sets of tangential double keys are used to secure the fly-wheel to the 
shaft. The only feature of the connecting rod requiring special notice is the construc- 



Fig. 612 —Olds Type “ K ” Engine. 


tion of the ends. The bearing metals are in detachable shells at both the crank and 
piston ends. 

As the trunk pistons of the large engines present an extensive rubbing surface, 
•the internal friction from this source is very great on standard designs. The pistons 
of these particular engines, besides being equipped with seven regular piston rings, are 
provided with a number of dove-tailed grooves, filled with white, anti-friction metal. 
This material is thoroughly peened into place, and then machined to size. The piston 
itself is slightly smaller than the cylinder bore, so that the bearing comes on the 
alloy rings. The peculiar virtue of this metal is that it has all the anti-friction 
qualities of Babbitt, and at the same time is capable of withstanding the heat 
encountered in the gas-engine cylinder. In this way, the internal friction of the Olds 
engine is cut down to a minimum. 

By reference to the general view, Fig. 612, it will be seen that the valve mechan¬ 
ism is driven from a lay shaft, which in turn is operated from the main shaft by 
means of Brown & Sharpe hardened helical gears. The governor has a positive drive 
from this same shaft, through similar gears. Both inlet and exhaust valves are 
operated by the same cam. 



STATIONARY ENGINES 


453 





Fig. 613—Olds Type “ K ” Engine. 































































454 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


By referring to Figs. 614 and 615, the valve mechanism between the cam and 
the two main valves will be readily understood. The exhaust valve is lifted through 

the rock-lever, A. The inlet valve is opened by means of the linkage B and C, and 

a rock-lever E. This lever has no stationary pivot, but rocks on the block H. The 
governor linkage is very simple. The centrifugal action of the governor weights tends 
to work the vertical lever G away from or towards the engine, according as the speed 
drops or increases; and this lever draws with it the fulcrum block H. Thus, when 

increasing load tends to lower the speed, the action of the governor draws the fulcrum 

block H away from the center of the engine, giving the inlet valve an increased lift. 
The gas is admitted through the passage M, while the air enters through N. The 
throttle valves in N and M admit of hand regulation of the mixture to suit the gas. 
The sliding sleeve K absolutely prevents the mingling of gas and air between strokes 
when the inlet valve is closed. The main valve springs are located outside of the 

bonnets and are instantly accessible for adjustment. Each valve stem is kept in align¬ 

ment by two bushings well separated. 

The starting valve is illustrated in Fig. 615. At A it is shown removed from 

the cylinder head; and at B, it is shown in‘position. The linkage for operating the 

valve can be readily traced. Referring to Fig. 614, the cam for operating the starting 
valve is shown at Q engaging a roller on the bell crank R. In Fig. 615 the letters 
Q and R are again used to designate these same parts. The method of connecting 
this bell crank to the valve is clear, without further lettering. In the view A, the 
starting handle is shown in position for operating the valve. The ordinary running 
position for this handle is shown by the dotted lines. When the handle is in this 

position, the weight of the horizontal arm of the lever R causes it to swing away 

from the cam, thus, entirely clearing it when the engine is running, 

When it is desired to start the engine, the fly-wheel is barred around until the 

piston is just starting on the explosion stroke. In this position, the point of the 

starting cam, Q, is opposite the roller on the arm R . The gas and air hand valves 
being opened, the starting air is turned on at the stop valve, and the starting handle 
S is thrown to position shown in A, Fig. 615. This opens the air valve, admitting 
the compressed air to the cylinder, and starting the engine. The cam will admit air 
at every explosion stroke until the engine takes up its cycle. This usually occurs on 
the third revolution. After this, the starting air is shut off at the tank and at S. 
When the engine acquires sufficient speed, the exhaust cam roller X is pulled over to 
running position; thus giving the engine full compression, when it is ready for the 
load. 

Ignition. The ignition system is of the make-and-break type, the current being 
furnished by a low tension Bosch magneto. The point of ignition can be easily 

adjusted during operation. Fig. 616 shows the method of operating the magneto and 
the position of the igniter block. 

Lubrication. The main bearings on the standard engines are lubricated by ring 
oilers, the reservoirs of which are in the engine frame. All of these reservoirs are 
equipped with sight glasses to indicate the level of the oil. The crank pin is oiled 
from a collecting ring secured to the outside of the right hand crank cheek. The oil 
passes from this ring to the surface of the crank, through a hole drilled in the crank 

itself. Oil drops into the ring from the spout of a stationary sight-feed oiler. A sight- 

feed oiler, D, Fig. 613, supplies a wiper E from which the piston pin is lubricated, 
lhe piston and cylinder are lubricated by a sight-feed force pump, driven by an 

eccentric from the lay-shaft. The governor and main lay-shaft gears run in an oil 


STATIONARY ENGINES 


453 



Fig. 614 .— Olds Type “K ” Engine. 



































































































































































































































456 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Fig. 615.—Olds Type “ K ” Engine. 









































































































































































































STATIONARY ENGINES 


457 


bath. The hand-pump oiler, V, shown in Fig. 615 is used for lubricating the exhaust 
valve stem, through the oil pipe marked V in Fig. 614. In the same view the letter 
T indicates the oiler for lubricating the sleeve K. The oil supplied by the cup W, 
Fig. 614, works down the inlet stem, oiling the upper bushings, the rocker-arm yoke 
and the lower bushing in turn. Table 103 gives the principal dimensions of Olds 
Type “ K ” engines. 

Olds Gasoline Engines. The Olds line of internal-combustion motors includes, 
besides the producer-gas machines just described, gasoline engines of from 8 to 50 H.P. 
They also manufacture a complete line of agricultural motors of from 1£ to 12 H.P. 
Both of these lines can be equipped with fittings for operating on city gas, or distil- 



Fig. 616.—Ignition Gear, Olds Type “ K ” Engine. 

late. Finally, this company makes a line of small kerosene engines of from 3 to 12 
H.P., which can be operated on gasoline, kerosene, or alcohol. So efficient is the mixer 
on this last type of motor that the company claims to be able to vaporize crude oil 
with it; but owing to the large amount of impurities in this fuel, does not recommend it. 

The Olds Type “ A ” engines are built for a variety of purposes, for farm use, 
for direct connection to pumps, hoists, etc., and for general power purposes. 

This motor is a horizontal, single-cylinder, four-stroke cycle machine, and for 
general utility is equipped with hit-and-miss regulation. 

All engines of this type are furnished water cooled, either by means of the 
customary circulating tank, or by the new hopper jacket system, but, when desired, 
the 1£ and 3 H.P. sizes are also furnished air cooled for .such special work as fruit 
spraying and the like. 

The jump-spark ignition system is used on the entire Type “A” line. 

Instead of casting the frame in one piece, the Olds engines of this type have a 



Table 103 

OLDS GAS POWER COMPANY, 


458 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



ioeiqht in IOOOlbs. — AH dimensions qiven in inches. 
























































































































































































Table 103 ( continued ). 


STATIONARY ENGINES 


459 



|— 



— H 



~—H 



Note- t 0 s coeiqht in 1000 lbs.—AH dimensions qiven in inches 


































































































































































































































































460 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBTJSTION ENGINES 



Fig. 617.—Olds Type “ A ” Engine. 

fuel is securely sealed from the air by an easily 

The fuel mixer Fig. 618 is an application 
ingenious device, which not only atom¬ 
izes the fuel thoroughly, but also 
raises it from the main tank to the 
auxiliary reservoir. 

There are a large number of 
engines now on the market which 
dispense with the gasoline pump, but 
most of these do so merely by locating 
the mixer nozzle near the level of the 
main fuel tank. When the tank is 
full, the mixer tends to flood, and 
when the level is low, the mixture 
is too lean. The Olds mixer, on the 
contrary, raises the fuel to an auxiliary 
reservoir, from which it feeds at a 
constant head, no matter w T hat the 
tank level may be. 

The method of operation in this 
mixer will be apparent from an in¬ 
spection of Fig. 618. A is the mixer 
tube, which delivers the mixture to 
the engine at B, drawing air from 
the outside through the opening C. 

The fuel is entrained from the nozzle 
D in the blast of inflowing air. Owing 
to the restricted area of the Venturi 
tube at E, the velocity of the air 
drawn into the cylinder by the engine 
piston is very high. This atomizes 
the gasoline issuing from the nozzle 


cast iron sub-base, entirely 
separate from the engine bed 
proper, Fig. 617.. This sub¬ 
base, being of box pattern, is 
utilized as a fuel reservoir. 
As it is made of cast iron of 
sufficient section to make a 
rigid base, it forms a sub¬ 
stantial tank as well. In this 
way the hopper-jacket engine 
is completely self contained. 
When it is desired to mount 
such an engine on skids, there 
are but two parts to mount, the 
engine and the battery boX; the 
fuel and cooling tanks are incor¬ 
porated in the engine itself. The 
accessible screw' cap, as in other makes, 
of the Venturi tube. This is a rather 


B 




A 



PL 9 390 . 

Fig. 618.—Olds Carburetor. 




























STATIONARY ENGINES 


461 


very thoroughly. The fuel always stands flush with the nozzle—any excess in the 
auxiliary reservoir will return to the main , tank through the overflow pipe F. 

On the suction stroke of the engine,, the vacuum created in the mixing tube 
extends to the auxiliary reservoir, through the passage G. Both the pipes F and H 
being open, the atmospheric pressure in the main tank forces the liquid up the pipe 
H, filling the auxiliary reservoir. The reservoir cannot be flooded by the oil passing 
up the overflow pipe, as the pressure of the oil on the disk valve K raises the same, 
closing this passage during the suction stroke. 



Fig. 619—Olds Type “ A” Engine. 


At the end of the suction stroke, this valve will drop, however, allowing any 
excess in the auxiliary reservoir to return to the main tank. Thus, an absolutely 
uniform action is secured, no matter what the heights of the supply in the mam tank. 

The details of the construction in general and of the valve mechanism can best 
be understood by reference to Figs. 619, 620, and 621. From these it is evident 
that the piston, connecting rod, engine bed, etc. are designed in accordance with well 


tripH cf n n rl <3 rr <5 

The “ two-to-one ” gears are of the spur type, the teeth being machine cut, and 
the pinion being of bronze. The exhaust valve is operated from a cam which is 
secured directly to the driven gear, thus eliminating entirely the cam shaft and its 
torsional strain. The motion is transmitted to the exhaust valve - , ig- > 1 u 

a double bell crank B and a “pull-rod” D. The advantages claimed for this arrange¬ 
ment are that the gearing is enclosed in the bed casting, thus being protected from 


















































462 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION-ENGINES 


dust and grit; also, the larger gear, Fig. 622, touches an oil bath in the bottom of 
the bed, insuring lubrication. Furthermore, the link, or “pull-rod,” between the 
valve and the cam is, by virtue of this arrangement, in tension instead of compression, 



and can therefore be made very light. Finally, this rod is provided with a simple 
adjustment, consisting merely of two nuts at the end, for taking up wear. This one 
adjustment takes .up the wear in the entire mechanism, and is instantly accessible. 
















































































































STATIONARY ENGINES 


463 



PL.^94 . 


Fig. 621.—Olds Type “ A ” Engine. 

























































































464 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 

The jackets of all these engines are made in a separate casting from the cylinder. 
Thus, the entire engine is not disabled in case of accidental freezing of the jacket 
water. The jacket alone is broken, and this can readily be replaced at a nominal 
sum. On all engines where the jacket is cast integral with the cylinder, such an 
accident entails the replacement of the entire cylinder, etc. 

As stated before, Olds Type “A” engines are governed on the hit-and-miss 
principle. The inertia governor, located in the fly-wheel, is shown in detail in Fig. 
622. It operates by interposing a pidk blade, when the speed becomes too high, which 
prevents the bell crank C, shown in dotted line, from returning to its normal position. 
This blocks open the exhaust valve and causes a miss stroke. At the same time a 
rod passing from the exhaust-valve gear to the inlet valve, indicated in Figs. 619 and 
621, puts an additional tension on the inlet-valve spring, and thus prevents the 
opening of that valve. 

The following table shows the principal dimensions of Type “A” engines: 


Table 104 

OLDS GAS POWER COMPANY, 

Tyre: A Engines . 



u 

Si 

U: 

3: 

t 

0; 

it 

All Dimensions <S'ven In Inches. 

Pulleys. 

(5 end ft al. Dimensions . 

Engine Base. 

Base. 

Anchor doll. 

Wheels 

Stand. 

Smau'st. 

A 

B 

£ 

D 

E 

F 

6 

H 

I 

J 

K 

L 

M 

N 

O 

P 

Q 


a 

T 

U 

V 

w 

X 

Y 

z 

a 

b' 

Oia. 

Fce. 

Oia. 

Flu 

2A 

3 

500 

/5k 

24k 

4oi 

Elk 

lok 

b/k 

% 

ZSg 

- 

4 

5 1 

22 

II 

5 

20i 

4k 


2& 

14 

81 

8 


12 k 

% 

2 

/b 

28 

2 

(o 

4 



3A 

4k 

450 

nk 

3ik 

45,t 

244 

rnk 


6$ 

27i 

- 

lk 

64 

25k 

A3 

5k 

24 

ilk 

4k 

32 

/6k 

4 


30k 

/s 

2 

2 

lb 

28 

2j 

8 

5 

b 

4 

4A 


455 

H 

35k 

44 

al 

/3k 


7/f 

27a 

3/1 

2 

74 

27 

14 

/3 

Ofl 

zsk 

izi 

>0% 

34 

!8k 

/O 


32j 

!6% 


zk 

20 

33 

z% 

to 

£ 



5A 

a 

400 

zok 

38k 

M 


I3l 

7/1 

74 

3A| 

3,1 

2 

6g> 

28l 

i4 

(o 

27 

izi 

/o .| 

3 hi 

18% 

U 

7a 


/7 


2k 

20 

3b 

3 

12 

b 



6A 

12 

380 

2!k 

m 

634 

334 

174 

U>1 

Hz 

3Ji 

3* 

zk 

7 

35k\ /& 

& 

33k 

tb 

13 

43k 

23 

7 

/2 

41k 

2/ 

% 

2 k 

28 

40 

34 

/& 

6 













































































































































I 


STATIONARY ENGINES 


465 


Table 105 

SUMMARY OF' RESULTS 



October 10-11. 

October 11-12. 

October 12-13. 

Duration of Test. 

Hr. Min. 

Hr. Min. 

Hr. Min. 

Hr. Min. 

Hr. Min. 

Hr. Min. 

Engine running. 

8 9 

7 53 

10 38 

8 1 

8 2 

10 2 

Fires banked. 

15 56 

16 

13 18 

16 

16 

14 2 

Fuel. 







Coal used in running, lbs. 

721.75 

704.50 

845.50 

725.50 

638.25 

839.25 

Coal used in banking, lbs. 

14.50 

36.50 

44.00 

11.50 

14.00 

14.00 

Total coal-lbs. 

736.25 

741.00 

889.50 

737.00 

672.25 

853.25 

Salvage from screening, lbs. 

239.00 

266.50 

368.00 

311.50 

185.00 

202.00 

Salvage per cent of coal. 

39.50 

26.00 

41.50 

42.10 

28.20 

23.60 

Coal per hour of running, lbs. 

88.60 

89.40 

79.50 

90.50 

81.90 

81.20 

Coal per hour of banking, lbs. 

0.91 

2.28 

3.35 

0.66 

0.88 

1 .00 

Water. 







Water used in compressing air, both units, gals. 

3 600.00 

6 000.00 

3 840.00 

Water used in blowing hot both units, gals. 

8 200.00 

11 600.00 

9 800.00 

Time of blowing hot, minutes. 

1 

7.50 

24.50 

20.50 

Water used in each vaporizer, gals, per hour . . . 


2.70 


2.70 

2 

.70 

Water in each vaporizer, per lb. of coal 







burned, lbs. 


0.27 


0.25 


0.26 

Water used in each scrubber, gals, per hour . . . . 

105.00 

105.00 

105.00 

Total water used in both scrubbers, gals. 

1 900.00 

1 680.00 

1 960.00 

Water used in scrubber, per H.P. of producer, 







pounds per hour. 


7.30 


7.30 


7.30 

Water used in cooling each engine, gals, per min. 

23.00 

23.00 

23.00 

Total water used in cooling engines, gals. 

25 000.00 

22 100.00 

25 700.00 

Per cent of water pumped. 


1 .80 


1.80 


1.90 

Total water used in plant, gals. 

38 800.00 

41 400.00 

41 500.00 

Per cent of water pumped. 


2.90 


3.40 


3.10 


October 10. 


October 11. 


October 12. 


Pumps. 

Revolutions. 

Revolutions per minute. 

Discharge per revolution by displacement., gals. 

Slip determined by preliminary test, %. 

Actual discharge per revolution, gals. 

Discharge per minute, gals. 

Discharge per min., less water used in cooling, 

24.80 gals. 

Total gallons pumped to reservoir. 

Pressures 

Force main, ft. 

Suction lift, ft. 

Total. 


No. 1. 

15 652.00 
32.00 

40.62 

6.20 
38 10 

1 219.20 

No. 2. 

15 417.00 
33.13 
40.62 

1.70 

39.93 

1 322.88 

No. 1. 

19 440.00 
30.47 

No. 2. 

15 768.00 
32.78 

No. 1. 

15 668.00 
32.51 

No. 2. 

19 676.00 
32.69 









1 160.91 

1 308.91 

1 238.63 

1 305.31 

1 194.40 
584 062.00 

1 298.08 
613 992.00 

1 136.11 
724 838.00 

1 284.11 

617 657.00 

1 213.83 
585 066.00 

1 280.50 
770 867.00 

No. 1. 
164.80 

12 69 

No. 2. 

168.20 

11.90 

No. 1. 
166.00 
11.65 

No. 1. 
167.00 
11.19 

No. 1. 
166.70 
13.42 

No. 2. 
166.70 
12.67 

177.49 

180.10 

177.65 

178.19 

180.12 

179.37 


Duty. 

In million foot-pounds, based on plunger 


No. 1. 


No. 2. 


No. 1. 


No. 2. 


No. 1. 


No. 2. 


Average. 


displacement, less slip and cooling water. 


Per 100 lbs. coal fired during run. 

Per 100 lbs. of coal, including banking. 

Per 1 000 000 B.T.U. in coal fired during run 
Per 1 000 000 B.T.U., including banking 


119.78 

117.43 

96.60 

94.70 


130.90 

124.46 

105.57 

100.71 


126.26 
120.01 
101.82 
96.78 


126.44 
124.47 
101.97 
100.71 


133.52 

130.74 

107.68 

105.44 


137.41 

135.15 

110.80 

109.00 


129.05 

125.38 

104.08 

101.20 



































































































466 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 

Operating Results. The only figures available were obtained on tests of a pumping 
station at St. Stephen, N. B. This plant is equipped as follows: 

Two Pintsch suction-gas producers, 135 H.P. each, 

Two single-acting 4-cycle Type “ K ” engines, 125 H.P., 22" dia., 28' 
stroke, 150 r.p.m., 

Two-direct connected triplex pumps, capacity 1250 gal. per min., 13" dia. X12" 
stroke, geared to run at 32 r.p.m. 

The auxiliary apparatus consists of a compressor, driven by Pelton wheel and 
furnishing air for starting engines and producers. 

The tests were made Oct. 10, 11, and 12, 1907. The following table contains the 
principal results. Indicator cards were apparently not taken. 

The following are some additional figures contained in the report. 

Coal used.Scotch Anthracite, Chestnut 


Composition of coal, Moisture, %. 1-17 

Yol. matter, %. 7.60 

Fixed C., % .‘ 83.64 

Ash, %... 7.56 

B.T.U. per lb. of coal. 12 400 

Coal used per water H.P., lbs.,. 1.53 

Coal used per B.H.P. (80% pump eff.) lbs. ... 1.22 

Thermal efficiency on water lifted, % . 13.4 

Thermal efficiency on B.H.P., %. 16.7 

Temp, jacket water, inlet, °F. 44 

Temp, jacket water, outlet, °F. 72 

Jacket water per min., gals. 23.0 

Heat loss in jacket water in % of heat in coal. 37.0 


It should be noted in connection with the thermal efficiency figures, that the 
plant was running at only about two-thirds capacity. The amount of jacket water is 
excessive, due to the low outlet temperature. 

Olds Suction-Gas Producer. The Olds Gas Power Company owns the exclusive 
patent rights for the United States on the suction-gas producers of the Julius Pintsch 
Company, of Berlin, Germany. Fig. 623 is a sectional view of one of the 300 H.P. 
plants. The principal elements of the plant will be readily recognized; A being the 
generator, B the vaporizer, C the wet scrubber, and J the dry cleaners. 

As in all suction types, the entire system operates under slightly less than 
atmospheric pressure; the engine drawing gas from the main, or exhausters, creating 
the draught through the apparatus. The cross-section area of the generator is espe¬ 
cially large per rated horse-power, to insure low reaction temperatures, at rated loads, 
and consequent freedom from clinker troubles. 

In addition, these producers are provided with a cone-shaped grate, on roller or 
ball bearings, which may not only be rotated, but also allows the fire to be reached 
by pokers through the bars. Four poke holes are provided just above the grate ring. 
These are of such size as to give ample freedom for the stoking bar, but still cut the 
heat loss by radiation to a minimum. When especially poor coal is used, additional 
facilities for barring are provided in the top of the producer. This top is not integral 
with the producer shell, but rotates on ball, or roller, bearings indicated at F. Thus, 
the barring hole G, and the coal hopper D, may be swung to any desired position. 














STATIONARY ENGINES 


467 


The top plate and charging hopper are made gas tight by water seals. The gas leaves 
the producer by the pipe E, passes down through the vaporizer, and up through the 



scrubber. It reaches the two cleaners, J , through the pipe H. 
in the same shell as the scrubber, yet they are entirely £ 
other by the partitions MM. The cleaners are furnished 


While these are contained 
leparated from it and each 
in duplicate to permit the 





























































































































































468 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 

cleaning of one while the other remains in operation. The exit of the gas from the 
cleaners is shown at the back. 

When it is desired to isolate the scrubber, etc., from the generator or producer, 
the bell K is raised by means of the rack and pinion. This action first closes the 
opening into the scrubber, and then only opens the vent-stack pipe. The entire nest 
of vaporizer tubes may be cleaned in place, or entirely removed from the cleaner, by 
loosening two flanges, and without breaking joints in connections between the producer, 
vaporizer and scrubber. 

Perhaps the strongest feature about this producer is the method of securing a 
uniform mixture of steam and air for the fire. Steam from the vaporizer passes 


Fig. 624.—Front View of Riverside Gas Engine at the Watson-Stilhnan Plant, Aldine, N. J. 

through the pipe N into the jet blower 0; this blower, once set by hand, will deliver 
an absolutely uniform mixture of steam and air. The heated air is drawn from the 
producer jacket P, and is delivered with the steam, partly to the producer through Q, 
while any excess escapes through R. The jet blower supplies more air and steam than 
the gasification process requires, and as the producer draws its entire air supply from 
Q, it is obvious that whether the load be heavy or light, it will always get the same 
ratio of air and steam as long as the adjustment of the blower is unchanged. 

The water regulator S serves to maintain a constant level of water in the 
vaporizer, irrespective of steam consumption, and without waste. This regulator is 
connected with the vaporizer by a feed pipe delivering the water into the lower part of 
the vaporizer, and by another pipe connecting with the steam chamber of the vaporizer B. 











STATIONARY ENGINES 


469 


The inlet and outlets of the producer, vaporizer, scrubber and cleaner are con¬ 
nected through piping with a series of water gauges on the gauge board T. Any 
irregularity in operation due to clogging, leaks, etc., in above apparatus can be 
immediately discovered by observing the vacuums shown by the gauges. Steam 
pressure in the vaporizer is indicated by a suitable gauge on the board. 

Besides the engines so far discussed, the American market offers a number of 
other perhaps equally important machines, but owing to the comparative youth of the 
industry in this country, little information is available regarding their construction and 
next to none concerning their actual performance. It will therefore be possible only 
to give brief mention to any of these engines and to allow the reader himself to study 
the construction as far as that may be done from the pictures and cuts that were 
available. The information given was in some cases furnished in the shape of specifi¬ 
cations by the engine company concerned, but in the majority of the cases is due to 
the columns of Power, and the Report of the Committee of the National Electric 
Light Association. 

6. The Riverside Engine Co., Oil City, Pa. 

Figs. 624, 625 1 , and 626 1 , together with the following specifications, will serve to 
show that the construction of this engine is unique in many, respects. The company 
guarantees that the heat consumption of these engines on producer gas will not 
exceed 10 000 to 11 500 B.T.U. per B.H.P. hr., that the speed variation will not exceed 
2% up to capacity, and that they will carry momentarily over loads of 10%. 


SPECIFICATION OF RIVERSIDE HEAVY DUTY DOUBLE-ACTING TANDEM CLASS 

GAS ENGINE 


<F” 


Cylinders and Sole Plate. The cylinders are cast in three pieces, two of which are alike, 
f cast iron, l 

are of \ semi-steel, [ according to size, and comprise the heads, valve chambers, rod housings, and 
L or steel ■* 

combustion chambers. The third piece forms the cylinder barrel and jacket and is of air-furnace 
iron cast on end with a large sink head or riser. This piece, the most important part of a gas engine, 
is of absolute symmetry and extreme simplicity. Allowance is made for reboring the cylinder 
barrel, and this whole section can be as cheaply and more easily renewed than an independent 
cylinder liner. The three parts are fitted together with circular tongues and grooves, and held 
together and to the main frame by four steel tie bars, which take the tension strains. 

The cylinders are mounted on a heavy cast-iron sole plate having a machined top surface, 

which keeps the cylinders in alignment and permits of their free expansion and contraction. By 
removing the distance pieces, the cylinder parts can be slid endwise on the sole plate, giving easy 
access to the interior of the cylinder, piston, and piston rings. The rings can be cleaned or changed, 
without disturbing the piston or rod. 

The sole plate makes a perfect drip pan under all cylinders, keeping all oil drip from the 
foundation. The exhaust and inlet piping is attached to the sole plate, hence no piping except 
the water-jacket piping, has to be disturbed when the cylinders are moved. There is no overhead 

piping or wiring to interfere with a traveling crane . . . , 

The sole plate contains a duct or passage for delivering fuel to each inlet valve and is tapped 

at exhaust passage for the connection of the exhaust piping. 

Valves. Both inlet and exhaust valves are of the semi-balanced, water-cooled, poppet type, 
operated in a vertical position. All water to the cylinder jacket passes through the valves first, 
positively cooling them without any attention whatever. 


1 Power, May 5, 1908. 






470 
























































































































































































































































































































































































































































































































































































































































STATIONARY ENGINES 


471 


The rise in temperature of the water in passing through the valves is only a few degrees,with 
the result that the valves are always cool, are not warped out of shape, are easily lubricated, never 
stick, and wear as well as the valves of a steam engine. The seats take a polish and run without 
regrinding for years. 

The balancing pistons run in renewable liners and are lubricated by force feeds. These pistons 
provide a very large guide surface, assuring positive alignment for these valves, indefinitely. The 
valve seats are reversible, are easily renewed, and are located slightly below the bottom of the 
cylinder bore, so that any foreign substances are swept from the cylinder at each exhaust stroke. 
All valves are readily removed from the top of the cylinder without disturbing the cam shaft. 

Piston and Piston Rod. Each piston and a section of rod form one solid piece. An 
enlarged end of the rod forms a shouldered steel center around which the piston is poured and 
to which it is welded. The rod end has rough grooves turned on it, making an intimate contact 
between the rod and piston. This construction also provides a steel center into which the adjacent 
section of piston rod is screwed. A locked-jam nut which is flush with the piston-face holds the 
sections securely together when they are in place. 

The pistons and rod are supported at three points, viz., at the cross-head, the center rod 
support and at the end of the tail rod. These supports are fitted with adjustable babbitted 
shoes. 

The construction of the Riverside piston rod is such that a piston and rod section in either 


cylinder can be removed, without disturbing the other cylinder or the connecting-rod, cross-head, 
or any part of the valve gear. Piston rods are made from steel forgings with water passages 
machined from the solid forging. 

Piston and piston rods are water cooled, the water entering through a telescopic joint 
connected to the side of the cross-head. Circulation through each piston is positive and the 
overflow is so arranged that pistons are kept full of water. Water passages through the pistons 
and rod are large and easy, so that not over 10 lbs. pressure is required for circulation. The 
overflow is visible, hence the water cannot come to a boiling-point without attracting the engineer s 
attention. 

Piston Rod Packing. Metallic packing is used in all piston-rod boxes. The packing rings 
consist of segmental cast-iron rings carefully turned all over and held by spring rings surrounding 
them. The ring sections are doweled in position and two completely assembled packing rings 
occupy the same groove in the gland. The gland is split in the direction of its axis and the halves 
held together by countersunk screws. The whole packing may be easily assembled around the rod 
and then pushed into the box and fastened. 

Main Frame. The main frame or bed is of the heavy-duty, rolling mill type, with bored 
guides and main bearings cast in' a single piece, and is of great weight and extreme rigidity 1 he 
main bearing jaws are carefully machined to receive the main bearings, which consist ot Babbitt- 

lined shells according to size, These are backed up for their entire length by 

adiusting wedges. The shells can be removed by raising the shaft slightly, and are adjustable for 
wear in "both directions. The heavy main bearing cap locks across the top of main bearing jaws, 
and is provided with a large opening in its center for inspection. , 

The cylinder end of the main frame is squared and machined to fit the forward end of the 
cylinder, and has machined holes for receiving the tie bars which attach the cylinders 

The design of this frame, with its bored guideway, is such that a large part of the metal is 

flhovp the center line. m&kinff it exceptionally stiff. , , 

Cross-head. The cross-head is a heavy steel casting of the box type, fitted with adjusta e 
top and bottom shoes, which are carefully turned to fit the bored gmdeway The^ cross- ead pm 
is carefullv turned to size and fitted in the cross-head with a taper at both ends, t 
up and held bv bolts passing through a holding plate which covers the end of the p i . 

Two kevwavs are machined in the pin at 90° from each other. A single key holds the pin 
fast, relative to > the cross-head, and the two keyways allow the pin to be rotated one quarter 

turn and so distribute the wear evenly over the surface. , , f t g0 ag not to 

The shoes are hung on pins, within eccentric bushings, and cannot be set out so as not to 
bear uniformly for their entire length. The crank-pin boxes are heavy steel shells, lined w.th 

genuine Babbitt The cross-head boxes are of heavy phosphor bionze. . 

Crank-Mt The shaft is of the side crank built-up type, and is machined all over 


472 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


with the crank pin cast integral, pressed onto the main shaft by hydraulic pressure and doubly 
keyed. The crank-shaft is guaranteed against breakage during the life of the engine. 

Fly-wheel. The fly-wheel is of proper weight, with the face and edges of the rim turned 
to run true, and barring-over holes are provided. 

The wheel is made in complete halves, and held together at the rim by heavy steel “T” 
links shrunk in the side and by bolts through lugs on the inner side of the rim. The wheel 
halves are held together at the hub by heavy steel bolts and the wheel is keyed rigidly to the 
shaft. 

Valve Gear. The valve gear of the Riverside engine is the simplest ever applied to a 
multiple-cylinder engine, and consists of a single shaft mounted on top of the engine, running in 
self-oiling bearings. This shaft runs at one-half the speed of the main shaft, and carries the 
inlet and exhaust cams, cams for operating the oil pumps, and the timer for the ignition 
system. Power is transmitted to the inlet and exhaust valves by an inlet and exhaust lever 
hung on a single pin. All cams are keyed rigidly to the cam shaft. This construction makes a 
minimum number of joints subject to wear, although ample adjustments are provided for taking 
up any wear. 

Governing. The speed of the engine will be maintained uniform within 2% either way 
from the mean under normal operating conditions, but might momentarily exceed this should 
full load be thrown on or off instantaneously. In no case will the speed exceed a safe limit. 
The speed is controlled by a sensitive carefully balanced enclosed type self-oiling centrifugal 
governor operating a double-beat balanced disk valve, which throttles the mixture, thus varying 
the mean effective pressure to suit the load and giving a uniform number of impulses regardless 
of the load. The mixture of air and gas can be varied at will and is controlled from one point 
so that all cylinders get an absolutely uniform mixture. This method of governing results in as 
uniform speed as in any system, but is without the complexity and necessitates none of the 
necessary accurate adjustments at each inlet valve which are incidental to a cut-off-gear. 

Lubrication. The cylinders, valves, and stuffing-boxes are supplied with oil from force-feed 
lubricators. Lubrication to the cylinders is timed so that the oil is fed between the snap rings 
on the piston; thus the oil is swept over the entire surface of the cylinders and not fed into 
the cylinder at such a time as to burn up and do no work. Sight-feed oilers are provided for 
all other bearings except the cam shaft which is lubricated with ring oilers. 

Ignition. Ignition is by an improved system, consisting of magnetically operated spark 
plugs in each cylinder. The sparking points are in series with the magnet coils which produce 
the break. This gives a positive external indication by the movement of the armature as to 
whether the spark takes place within the cylinder. The inductive resistance of the coils 
augments the strength of the spark which is extremely long and hot. 

The timing is done electrically. No gearing or mechanical trips are used. Well insulated 
stranded wires, running through iron-armored conduits, lead from the timer to each spark plug. 

The timer is mounted on the cam shaft, and built in a heavy substantial manner. The 
contacts are of the wipe type, are made of tool steel, and are adjustable for wear. A visible 
indicating spindle shows the amount of each contact. 

The entire wearing parts of the timer run in oil, which prevents the burning of the contacts 
and reduces wear to a minimum. 

Wiring from the timer to each spark plug is protected by a five-ampere, enclosed fuse; 
thus, should any spark plug become damaged or short-circuited, the other plugs would not be 
affected. 

Any spark plug can be removed and replaced while the engine is in operation. 

B\ individual adjustments, the time of ignition in any cylinder end can be set; and by a 
single handle, the time of ignition in all cylinders can be varied simultaneously. Any of these 
adjustments can be made equally as well with the engine in operation as otherwise. 

Pneumatic Starting Device and Control. The pneumatic starting gear consists of two 
shifter pistons mounted within the valve lever carrying pins on the rear cylinder. These pistons 
are shifted by throwing a small three-way cock, which applies compressed air to the one side 
ot the piston, which in turn moves the valve levers, bringing the cam rollers in line with an 
auxiliary cam. This puts all the valves in both ends of the cylinder into two-cycle action- or 
in other words, makes a poppet valve air or steam engine out of this cylinder 

This permits the engine to be started on any stroke and on either quarter. Since the 
operation is entirely automatic, the engine will run as long as compressed air is applied. 

I he other double-acting cylinder continues to operate as a 4-cycle gas engine, and takes up 


STATIONARY ENGINES 


473 


its explosions after the first revolution. Compressed air is then shut off from the starting 
cylinder, and the three-way cock reversed. This permits the cam levers to return to their 
normal position, and that cylinder immediately goes into 4-cycle action. 

The starting gear adds practically no complication to the engine, as the regular inlet and 
exhaust valves are used for distributing air. 

There are no extra shafts or auxiliary valves, nor any tappings into the cylinders. There is 
only one compressed air pipe leading to the sole plate, air being delivered into the fuel duct and 
entering the cylinders via the inlet valves. 

On a controlling pedestal is mounted a lever for controlling the free air supply, the three-way 
cock, the compressed air throttle wheel, the lever for controlling the gas supply, and the ignition 
switch. This permits the engineer to start the engine, bring it up to speed and adjust the air and 
gas to a correct mixture without leaving his position. 



Fig. 627,—4-cycle Double-acting Tandem Engine, Mesta Machine Co. 



Fig. 628. 


r—r™"" —- - -»- 


B.H.P. 






















































474 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


SPECIFICATIONS FOR MESTA GAS ENGINE 

The Mesta gas engine is a horizontal, double-acting 4-cycle machine built either in single- or 
twin-tandem units and follows what might be called the standard lines of heavy duty prime 
movers of this kind, laid down in the well-known Nurnberg type, modified to meet the 
American requirements. 

Frame. The main frame or bed plate is plain and massive in its construction, combining 
strength and accessibility. Special attention has been given to the rigid connection between the 
bed plate and forward cylinder obtained by a double flange construction connected by heavy 
ribs. Following the American practice the crank pin is overhung and the main bearing is 
designed to meet this arrangement. The bearing is made of four parts with liner adjustment 
for the forward quarter box and wedge for the rear. This method offers less possibility of 
improper alignment so often found where both quarter boxes are wedge adjusted. The bearing 
cap is of the interlocking construction, acting both as compression and tension member over the 
jaw of the bearing, thus insuring great rigidity. 

Cylinders. The cylinders are cast of vanadium steel in order to withstand the high vibratory 
stresses caused by alternate explosions in the two cylinder ends and also in order to reduce the 
section of cylinder walls to a minimum, thus increasing the cooling effect of the jacket. The 
central portion of the jacket is cast open and covered with a split band, which can easily be 
removed, in case it should be required for any extensive cleaning around the cylinder of accu¬ 
mulated sediment from the cooling water. The lower part of the cylinder is also provided with 
covers which permit an easy access to the annular space around the exhaust-valve bonnet. 

The inlet valve is located on top of the cylinder and the exhaust valve on bottom. This 
arrangement has the advantage of symmetrical structure, tending to equalize the strains in the 
cylinder. A further advantage in this arrangement of placing the inlet and exhaust valves as 
far apart as possible is that the heating of the incoming mixture is diminished. The heating 
results in decreased unit density of mixture, and therefore means loss in the capacity of the 
engine. In placing the exhaust valve at the lowest point in the cylinder, dust and foreign 

matter is more easily swept out than in any side chamber construction. 

Valves. Both inlet and exhaust, valves are of the mushroom type and operated by means 
of rolling levers and pull rods actuated by a single eccentric for each two valves. On the inlet 
valve stem is mounted a mixing valve consisting of a hollow sleeve which closely fits the 
annular gas and air passage. Ports in this sleeve register with ports in the gas and air passage, 
when the valve through its rolling lever is opened. The quality and quantity of the incoming 
mixture is controlled by independent valves in the gas and air passages actuated by the 

governor according to change of load. 

Pistons. An important feature in the design of this piston is its being cast in one piece, 
with a perfectly symmetrical distribution of metal, thus overcoming to a large extent the 

liability of shrinkage strains and resulting tendency to crack under service. The water 
circulation inside the piston is obtained without the use of adjustable parts. There are no cored 
holes in the outer walls and a smooth surface is presented without metal plugs and their 
liability to leak. The piston is held floating free from the cylinder wall by means of an inter¬ 
mediate and rear cross-head and the piston rings touching the walls are so designed as to give 
a uniform bearing pressure. 

The Piston Rod is made in halves and connected in the intermediate cross-head. Pro¬ 

vision is made for adjustment of the rod both in the horizontal as well as vertical plane. 
Water is introduced to the rod in the center and discharged at both ends. 

Cylinder Heads. These are cast in one piece, cored out for cooling water, and properly 
recessed for the metallic packing, which is set so far back from the inner end of the head as 
not to require cooling in addition to that furnished for the head. The construction of the head 
is such as to allow for changes in compression if desired. 

Ignition. Two igniters are placed in each end of the combustion chamber, one at the top 
and one at the bottom, and are of the magnetic make-and-break type. A timing device controls the 
spark, which can be advanced or retarded to suit condition. Provision is made for connecting 
the timer with the governor in order to be able to change the point of ignition for change in 
load. 


STATIONARY ENGINES 


475 


Air Starting. The simplest method of starting is obtained by introducing compressed air 
to the cylinders at a period corresponding to the power stroke in normal operation. This is 
accomplished by four cam operated poppet valves in the air main to the cylinders which auto¬ 
matically fall out of gear after the air is turned off. Check valves are located in the cylinder 
at the point of air admission and as soon as the engine gets an explosion the valve will close 
due to the higher internal pressure. The air is then turned off and the engine is now in 
operation. 

8. The Allis-Chalmers Co., Milwaukee, Wis. This company was originally the 
American licensee of the Niirnberg engine, but has within the last year or two 
developed an engine of somewhat different type. The main changes from the Niirn- 
berg' design apparently consist in the use of the side-crank frame and in the com¬ 
bining of the main inlet and mixing and regulating valve into one housing. Figs. 
629, 630', 631 and 632 will serve to show the important features of the design. 
The inlet-valve gearing and the regulation being radically different will need special 
mention, and the following description, in connection with Fig. 633, is taken from the 
Gas Engine Report above mentioned. 

“ The operation of gas valve will be readily understood by referring to the figure. 
At H is shown the gas valve. This valve is of the double-seated type and is located 
concentric with the main inlet-valve stem. The gas enters to the space above this 
valve from the gas and air manifold N. Valve H is operated by means of two rods 
which connect it with cross-head V, roller lever U and lever pin S, which pin is 
allowed to move up and down with the small cross-head V. The fulcrum lever T is 
forked on its inner end, and the ends of this fork are pivoted by pins to the valve 
bonnet; these pins being stationary in this bonnet and not moving with cross¬ 

head. 

Rolling lever U is connected to main inlet valve rolling lever through the 
connection as shown, and both levers move in unison. The end of fulcrum lever T 
is connected by means of its rod to the small eccentric placed on the shaft F. This 
shaft is rotated by the governor through the rod E, and this causes the end of lever T 

to take a different position dependent on the governor and thus the time of opening 

and lift of gas valve is changed accordingly. 

“It has been found that for gases such as blast furnace and producer gases, a 
double-seated valve will work freely, even after it has become coated with impurities 
contained in these gases, and it does not tend to stick, as would any type of 
cylindrical or piston-valve construction, and due to this it is possible to operate 

a double-seated valve through the means of a simple governor, and not necessary 
to complicate the governor arrangement by resorting to the use of a relay of any 
description. 

The small eccentric before mentioned is so arranged that it can be thrown out ol 
gear and held in such position that the gas valve has no lift. This is useful when it 
is desired to cut off the gas from any of the cylinder ends. After the gas passes the 
valve H it mixes with the air which passes in through the holes arranged in the 
sleeve extending into air passage; the flow of air is at nearly right angles to the 
current of inflowing gases, thus obtaining thorough mixing of gas and air on their 
passage to the cylinder. The air is proportioned by a valve located m air and gas 
manifold N: this valve being operated through means of the hand-wheel shown at D, 
and, when once set, does not need changing unless there is great variation in the 

quality of the gas supplied.” 




476 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Fig. 629.—Allis-Chalmers 4-cycle Double-acting Twin Tandem Engine. 



















































































































478 


Fig. 631.—Plan View, Allis-Chalmers Engine. 















































































































































































































































































































































































































































































STATIONARY ENGINES 


479 







Fig. 632.—End View, Allis-Chalmers Engine. 









































































































































































































480 CONSTRUCTION, ERECTION, TESTS OF INTERNALrCOMBUSTION ENGINES 




















































STATIONARY ENGINES 


481 


9. The William Tod Co., 
Youngstown, Ohio. The en¬ 
gines built by this company, like 
all of the large American engines 
with the exception of the De la 
Vergne-Korting, is of the 4-cycle 
double-acting tandem type. Figs. 
634, 635, and 636 show the general 
make-up of a 30X42" engine, 
which has been in service since 
March, 1908. A 42X60" engine 
direct-connected to two 80x60" 
blowing tubs is now building. 

The three general views above 
mentioned, together with the cross- 
sectional cuts in Fig. 637, will 
serve to explain the construction. 
The use of the side crank frame 
and the split cylinder jacket has 
about become standard practice. 
The cylinders are cut in half, 
bolted at the middle, and are 
supported at the ends only, the 
two C3’ r linders, the distance piece, 
the tail frame, and the main frame 
all being held together by four 
strong continuous tie-rods. 

The valves are located top 
and bottom. For one cylinder- 
end they are operated by one 
eccentric, the exhaust valve direct, 
the inlet valve by means of a 
beU crank. There is nothing out 
of the ordinary in the construc¬ 
tion of either exhaust or main 
inlet valve, but the mixing and 
governing arrangements are differ¬ 
ent from those found in any other 
American engine. 

The engine governs by propor¬ 
tioning the amount of gas to the 
load, there being practically con¬ 
stant compression. The mixing 
valve is a three-seated sleeve shown 
in Fig. 638 just above and concentric 
with the main inlet valve. This 
valve moves up and down with the 
inlet valve. Above this mixing 
valve is placed the regulating 



Fig. 634.—Double-acting Tandem 4-cycle Engine, Wm. Tod Co. 










482 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Fig. 635. —Tod Double-acting Tandem Engine. 







STATIONARY ENGINES 


483 



Fig. 636.—Tod Double-acting Tandem Engine. 

















4 






















































































































































































Fig. 638—Cross-section of Cylinder and Valves, Tod Double-acting Tandem Engine. 

























































































































































486 CONSTRUCTION. ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


valve of which a plan view is shown in Fig. 639. This valve consists of four disks 
having segmental slots as shown. The upper disk is oscillated about its center by 
means of the longitudinal rod H which passes along both cylinders and operates the 

disks in the other valve housing in exactly the 
same way. Owing to the motion, the slots in 
the disk alternately open and close similar slots 
in the stationary disk directly below it. Rod H 
receives its constant back-and-forth stroke from 
a small crank which is clearly shown above the 
cross-head guide in Figs. 634 and 635. At full 
load the motion of the small crank is in phase 
with the motion of the main crank, that is, 
the disk valve opens at the beginning of the 
suction stroke, and furnishes gas full stroke. 
As the load drops, the governor, by means of 
the floating spur gears shown in Fig. 640, causes 
the governor crank to lag behind the main 
crank, thus retarding the time of opening of 
the gas valve, the cylinder drawing only air 
during the first part of the stroke. 

This method of governing has been found entirely successful in practice as long 
as the quality of gas was fairly constant or did not vary too suddenly. To take care 
of a change in the quality of gas, there is placed under the stationary disk a disk 
throttle whose position may be changed by hand through a worm, thus adjusting the 
gas to air according to the quality of the former. 

No operating results from this engine are available. Although the heating value 
of the gas varied suddenly from 70 to 110 B.T.U., which caused some irregularity in 
firing, the engine regulated satisfactorily and carried overloads up to 20% for long 
periods. 

10. The A. H. Alberger Co., Buffalo, N. Y. This firm builds a type of engine 
which is finding much application for medium powers, that is, the single-acting tandem. 
The main features of the construction become plain from Figs. 641, 642 and 643. The 
design has the merit of great simplicity. An unusual feature in the cylinder construction 
is found in the use of an unjacketed head for the back cylinder. The valves are placed 
side by side in a chamber at the side of the cylinder, as shown in Figs. 641 and 644. 
Neither valve is water-cooled and both are mechanically operated from a lay shaft under¬ 
neath by means of simple cams. A small eccentric on the lay shaft, Fig. 641, also 
operates the make-and-break igniter at the inlet side of each valve chamber. The two 
inlet valve chambers are connected by a cast iron header, at the middle of which h 
mounted the mixing and governing valve, see Fig. 645. This valve consists of a 
hollow cylinder divided into two parts by a partition across it. Ports cut in the side 
of the cylinder match with similar ports in the side of the cage surrounding the valve. 
Gas is admitted to the header through the upper ports of valve and cage, and air through 
the lower. The vertical adjustment of the valve in the cage controls the ratio of air 
to gas. Raising the valve by means of a thumb screw located on the outside of the 
mixing chamber decreases the gas ports but increases the air ports by the same 
amount, so that the total effective port area remains the same. Lowering the valve 
has the opposite effect. The mixture is admitted to the header by oscillating the 
valve in its cage, alternately opening and closing the ports first for one cylinder and 



Fig. 639.—Governing Valve, Tod Engine. 

























Fig. 640.—Governor, Tod Engine. 







































































488 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Fig. 641. —Alberger Single-acting Tandem Engine. 



Fig. 642. —Cylinder Construction, Alberger Single-acting Tandem Engine. 


















STATIONARY ENGINES 


489 




Fig. 644.—Valve Chest, Alberger Engine. 





















































































































































490 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


then for the other. This motion is produced by connection to the pin of a Rites 
inertia governor shown in the wheel in Fig. 641. The position of the governor cuts 



Fig. 645. Governing Valve, Alberger Engine. 


the mixture off earlier or later during the suction stroke. The use of a governor of 
this type is unique in gas engine practice. The regulation with this system of 
governing can be made very close, but, as applied here, it is not possible to change 

speed while the engine is running. 
This is desirable sometimes in water¬ 
works engines, in engines driving 
air, gas, or ammonia compressors, 
etc., for which reason the firm also 
uses a throttling (fly-ball) governor, 
operated from the lay shaft, as shown 
in Fig. 646. 

Alberger (Buffalo) tandem gas 
engines are built in sizes from 35-250 
B.H.P. for producer gas and from 
45-300 B.H.P. for natural gas. Twin 
tandem units, the layout of one of 
which is shown in Fig. 643, double 
these capacities. 

11. The Struthers-Wells Co., Warren, Pa. Figs. 647, 648 and 649 show a 
type of tandem engine, built by this company, which is distinctly different from the 
one just described. The illustrations are sufficiently clear to explain the construction 
without further description. The construction of the frame, cylinder heads, and barrels 
valve cages, etc., with a view to accessibility and easy removal, is somewhat out of 
the ordinary and deserves special mention. 





Fig. 646.—Throttling Governor, Alberger Tandem Engine. 







































STATIONARY ENGINES 


491 


This firm also builds several other types of engines, both vertical and horizontal, 
but information regarding economy, etc., did not seem to be available either for these 




or for the engine above illustrated, 
tion states that the firm guarantees 


The Report of the National Electric Light Associa- 
for the tandem engine a gas consumption of not 


Fig. G48.—Cross-section, Single-acting Tandem Engine. 

















































































































492 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


over 10 cu.ft. per B.H.P. hour for natural gas of approximately 1000 B.T.U. per cu.ft 
It is also stated that on a recent test the consumption, was shown to be 8.1 cu.ft 
per B.H.P. hour, the gas containing a little less than 1000 B.T.U. per cu.ft. 



Fig. 649. —Cross-section of Cylinder and Valves, Struthers-Wells Single-acting Tandem Engine. 










































































































































PORTABLE AND SELF-PROPELLED ENGINES 


493 


B. PORTABLE AND SELF-PROPELLED ENGINES 

Under this head are considered simple portable power trucks, traction engines, 
automobile and marine engines. The various types built are nearly all operated with 
liquid fuel (gasoline, kerosene, and lately also alcohol), mainly because these fuels 
have the greatest heating value per unit of volume. 1 An exception to this rule is 
found only in gas traction engines or gas railways, in which the fuel consists of a 
comparatively small quantity of illuminating gas highly compressed (see p. 508). 


I. Portable Engines 

Fundamentally our present day portable oil engines are in reality nothing more 
than engines mounted on trucks, together with muffler, oil tank, water tank, and other 
auxiliaries. A construction which unites into one harmonious whole all of these 
component parts, such as is exhibited by portable steam engines, has not yet appeared 
and is in any case not as easy to obtain as in steam locomobiles, the main part of 
which consists of the boiler, which in turn may be made to carry the auxiliaries. 
The oil locomobile, on the other hand, requires a special frame to act as foundation 
for the engine and the only thing remaining to the designer then is to place the 
auxiliaries as best he may to obtain good appearance. 

The most important market for the portable oil engine is of course created by 
the agricultural demand, and builders have naturally paid most attention to the 
needs and requirements of this industry. The builder usually bases the power capacity 
of the locomobile under consideration upon the nominal power demand of his threshing 
machine, which requires most power among agricultural machines. Since the internal- 
combustion engine lacks the overload capacity that would be possessed by a steam 

locomobile, to be on the safe side the oil traction engine should be given a maximum 

capacity at least 40 or 50% above the normal power demand of the thresher. Failure 

to take this important point into account is always followed by trouble for both 

buyer and builder. 

The threshing drums usually make from 1200 to 1400 revolutions per minute. 
The diameter of the pulleys is from to 8" with a width of face of about 8". The 
belt speed is in the neighborhood of 40 ft. per sec., so that this as well as the ratio 
between driver and driven pulley is comparatively high. For good operation the 

distance between shafts should not be less than 15 ft., and the heavy belt not only 

brings considerable stress to bear on the crank but also renders difficult the starting 

of the larger sizes. This trouble is often avoided by the use of a countershaft which 

drives the drum, so that the engine may be started with the mam driving belt at 
rest. The same end is reached by simpler means by use of any of the well known 
designs of friction clutch. The belt is to be placed in such a way as not to interfere 
with the attendance of either engine or thresher. The fuel and water supply carried 
should last at least ten operating hours. 


1 A steam locomobile for each effective B.H.P. requires at least 6* lbs. of coal and about 45 lbs. 
of water, together, say, 50 lbs. of material. An oil locomobile requires only from about 7to.9 lb_o 
oil for fuel, and, if cooling by vaporization is used, about 2.2 lbs. of water per B.H.P. hour The 
weight of these materials therefore that must be carried per H.P. hour in the oil engme is therefore 
-only from to -£> of that required by the steam locomobile. 





494 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


The fuel consumption of locomobiles, as in general that of all portable and self- 
propelled motors, is usually somewhat greater than that of similar stationary engines 
of the same capacity. The main reasons for this are found in unavoidable jars and 
vibrations to which the vehicle is subject and which absorb a certain amount of the 
power developed, and the less perfect formation of mixture, cooling, etc. 

Designs of Locomobiles: 



Fig. 650. —Deutz Alcohol Locomobile. 


The portable alcohol engine of the Gasmotorenfabrik Deutz, during the last tests 
under the auspices of the German Agricultural Society, showed the best results. The 
fuel consumption varied as follows: 

Alcohol per B.H.P. hr., lbs., .802 at 16.8 H.P., maximum load. 

.855 at 12.1 H.P., normal load- 
1.113 at 6.27 H.P., half load. 

and 4.62 lbs. per hour at no load. 

Table 106 shows the dimensions of the usual sizes of locomobiles built by this 
company. 


Table 106 

DIMENSIONS AND WEIGHTS OF DEUTZ ALCOHOL LOCOMOBILES 


Normal rating, B.H.P. 

4 

6 

8 

12 

16 

20 

R.p.m. of driver on countershaft. 

450 

445 

420 

420 

375 

345 

Approximate weight, lbs. 

4620 

5720 

6820 

8800 

11450 

14300 

Maximum length, without tongue, ft. 

9.85 

10.18 

10.50 

11.15 

12.45 

13.10 

Maximum width, ft. 

Maximum height, ft. 

4.75 

5.08 

5.42 

6.08 

6.72 

7.20 

6.90 

7.05 

7.20 

7.85 

8.53 

9.00 

Diameter driving wheel, in. 

21.3 

21.3 

23.7 

23.7 

26.0 

28.4 

Face, driving wheel, in. 

5.5 

6.3 

7.1 

7.9 

8.7 

9.5 

Wheel base, in. 

59.0 

63.0 

67.0 

73.0 

82.7 

89.5 

Wheel tread, in. 

36.6 

39.8 

42.9 

48.5 

52.3 

57.5 

Width of tire, in. 

4.72 

5.12 

5.51 

5.92 

6.31 

6.70 

Diameter front wheels, in. 

28.2 

29.5 

29.5 

31.6 

35.4 

35.4 

Diameter rear wheels, in. 

39.4 

39.4 

39.4 

43.4 

47.3 

47.3 



























PORTABLE AND SELF-PROPELLED ENGINES 


495 







ro 

O' 



:. T 


~Fl 


Figs. 651-653.—“Gnom” Lo¬ 
comobile, 8 H.P., Motoren- 
fabrik Oberursel A.-G.,Ober- 
ursel, near Frankfurt M. 

Water is cooled by means of 
cooling tower. 

For details of engine see Figs. 

526, 527. 


Table 107 


DIMENSIONS AND WEIGHTS OF GNOM LOCOMOBILES 


Normal rating, B.H.P. 

2 

3 

4 

5 

6 

8 

10 

12 

15 

20 

R.p.m. of countershaft. 

360 

350 

300 

300 

250 

240 

250 

250 

250 

250 

Weight including counter,* lbs. . . . 











Diameter j of countershaft or \ in. . 

19.8 

19.8 

19.8 

23.6 

35.5 

35.5 

39.4 

39.4 

39.4 

39.4 

Width \ driving pulley, / in. . 

5.5 

6.3 

8.6 

8.6 

5.5 

5.9 

7.5 

7.5 

7.5 

9.1 

Width of belt, in. 

2.36 

2.76 

3.53 

3.94 

4.72 

5.12 

5.90 

6.70 

6.70 

7.90 


* Counter-shafts used above 6 H.P. The driving pulley or the counter-shaft runs a trifle slower than 
the pulley on the crank-shaft (see figures in Table 62, p. 361). 



Fig. 654.—Wenzel Locomo¬ 
bile, built byFr. Zimmer- 
mann & Co. Akt.-Ges., 
Halle S. 

Water cooled by extended 
surface radiators. 


Friction-transmission gear. 



























































































































































































































































































































496 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 





; * 

E o iut^ooj 


Table 108 

DIMENSIONS AND WEIGHTS OF ALTMANN KEROSENE LOCOMOBILES 


Rated B.H.P. 

2 

4-5 

6-8 

8-10 

10-12 

12-15 

16-20 

20-25 


220 

220 

220 

220 

220 

200 

200 

200 

Approx, weight, lbs. 

Max. length (without. 

2530 

4400 

5060 

6600 

7700 

8800 

4900 

11000 

tongue), ft. 

7.22 

10.65 

10.65 

10.65 

10.65 

11.50 

11.50 

11.50 

Maximum height, ft. 

3.60 

5.90 

5.90 

5.90 

5.90 

5.90 

5.90 

5.90 

Maximum width, ft. 

7.22 

7.22 

7.22 

7.22 

7.22 

8.20 

8.20 

8.20 

Fly-wheel f diam. and \ in.. . 

39.4X3 

47.2X3.9 

47.2X5.9 

51.2X7.9 

51.2X9.1 

59 X 9.9 

59 Xll 

59 Xll.8 

Belt pulley l width / in... 

13.8X5.9 

19.7X3.9 

19.7X3.9 19.7X7.1 

19.7X7.1 

23.6X7.9,23.6X11 

31.5X11.8 



































































































































































































































































































PORTABLE AND SELF-PROPELLED ENGINES 


497 



Fire Pumps operated by gasoline engines are superior to steam fire pumps in 
that they are ready for immediate operation and, on account of their smaller weight, 
are easier to handle. To do away with the use of teams these pumps have also 
been constructed as automobile pumps. 























































































































498 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


II. Motor Plows 

The oldest motor plow was built by Ganz & Co. and was introduced in 1896. 
This plow was direct-connected, that is, the implement itself was directly driven by 
the engine, and it represented the first useful solution of the long-standing problem. 



Ganz & Co., following the patents of their former general manager, Mr. Mechwart, do 
not use the ordinary plow share, but a sort of “ drum ” plow, illustrated in Figs. 
661-663. This is suspended at the rear of the locomobile and turns with moderate 
speed while the locomobile moves ahead. 1 A 12 H.P. gasoline engine used in the first 


1 For greater detail, see Z. d. V. D. I., 1896, p. 1367. 












































































































































499 


PORTABLE AND SELF-PROPELLED ENGINES 

types proved itself considerably too weak for rapid and sure operation in difficult 
soil. The power capacity was then increased step by step until the final motive power 
consists of a 40 H.P. four-cylinder vertical twin engine (each pair of cylinders in 
tandem) making 600 revolutions per minute. 

Of German gas-engine builders only the Motorenfabrik Oberursel builds motor- 
plows, Figs. 664 and 665. They pay special attention to the use of alcohol as fuel. 
The first attempts of this firm at building direct-acting motor plows do not seem to 
have been very successful, because the alcohol motor plow made by them to-day 
employs the so-called double machine system. This consists in placing a locomobile 
at each side of the field and having them draw a plow back and forth over the field 
between them by means of a wire rope. After each furrow the machines are moved 
sidewise the required distance. 

The construction of such plowing machines does not offer any particular difficulties; 
the arrangement of the drives, etc., may be directly based on the designs of successful 
steam plows. The main point again is to make certain that there is ample engine 
capacity. Less than 20 B.H.P. should never be used, and it is better to commence 
with from 30 to 40. But under such conditions the motor plow no longer possesses 
any advantage over the steam plow regarding operating costs. 

In spite of this fact agriculturists in general take a great deal of interest in the 
further development of the motor plow, especially with reference to the use of alcohol 
as a fuel. The principal thing expected from them is the general introduction of 
machine plowing which at present is not possible on account of the high cost of steam 
plows. The total capital cost of plowing machinery is of course least with the single 
machine system, for which reason, and in spite of the unfavorable results so far 
obtained with steam plow’s working on this system, attempts are constantly made to 
solve the problem with motor plows. It is true that the single machine system is 
best adapted to the motor plow because there is a pause of only about one minute, 
while the machine is changing to the next furrow, for every period of operation 
lasting from 3 to 5 minutes. The engine is therefore in use a large proportion of the 
time. In the usual two-machine system, on the other hand, each engine works for a 
period of from 3 to 5 minutes, plowing one furrow, and then rests for a period 
lasting from 5 to 8 minutes, while the opposite engine plows the next. Now 
it is well known that an oil engine is quite sensitive against such continued inter¬ 
ruptions in the operation and starts again only with comparative difficulty. Hence it 
becomes almost necessary to keep the engine in motion also during the idle periods, 
which of course seriously affects the economy of the entire system. The single 
machine system is therefore not only cheaper as to first cost, but also more economical 
in operation. 

As points in favor of the motor plow may be mentioned their lighter weight and 
the lower demand for fuel and water, as compared with steam plows. The fuel cost 
of moving them around from one place of operation to another is consequently less. 
Under certain circumstances these advantages may be more than sufficient to over¬ 
balance the greater costs of the kind of fuel used. Considerations of this kind of 
course by no means assure the future of the motor plow. The market for them is 
probably not extensive enough to lead manufacturers to undertake any extended and 
costly series of experiments, especially not at this time when the gas engine industry 
finds an apparently much more promising field in the building of blast-furnace gas 
engines, suction-gas plants, etc. 


500 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 
































































































































PORTABLE AND SELF-PROPELLED ENGINES 


501 


III. Motor Vehicles 

This term includes automobiles and automobile trucks, for the transportation of 
passengers and freight, and motor boats. In a wider sense it also comprises self- 
propelled fire engines, traction engines, etc. The latter, however, will be treated, 
together with motor railways, under the next heading. 

Concerning automobile engines proper, the author refers to his little special treatise 
on this type of engine. 1 In addition some of the later constructions are given below. 
The power required by an automobile may be estimated with sufficient accuracy by 
allowing 2.2 B.H.P. for every 1000 lbs. of total weight and for a speed of 10 miles 
per hour, assuming average roads and weather conditions. 2 It is not possible to give 
a generally applicable method of computing the power required on account of variation 
in the resistance encountered and the general uncertainty regarding them. Of the 

power developed in the engine, from 20 to 50% (and in especially bad cases even 
more) are lost in transmission between engine and driving wheels. The average 
powering of rtiotor vehicles shows a steady increase. At present light automobiles are 
supplied with from 6-10 H.P., the heavier cars have from 15-20 H.P., while some of 
the racing cars show from 60-80 H.P. 

All motors vehicles must be able to reverse their direction of motion (to back up). 
But since up to now no successful scheme has been worked out to reverse the engine 
itself, this can only be done by means of some auxiliary apparatus. Such change or 
reversing gears are interposed between the driving axle and the crank-shaft. In the 
case of motor boats, a device very often used is the reversible propeller in which, by 
varying the position of the blades, ,th e direction of motion of the boat may not only 
be reversed, but intermediate speeds either ahead or astern may be obtained. But 

all of these reversible propellers when used under considerable stress any length of 

time develop bad faults, so that their use for large powers is almost out of the 
question. As soon as the power exceeds 60 or 80 H.P. there is no doubt that the 
ordinary screw with solid blades will give better service, and motion astern will then 
have to be obtained by the use of some reversing gear or clutch. This does not 
mean to say that the latter scheme is perfect, but it is the most reliable as long as 
a good way of reversing the engine itself is as yet not available. Until the latter 
problem is solved it is quite likely that the use of internal-combustion engines in 
marine practice will be largely confined to the smaller capacities. 

In consideration of the fact that the crank-shaft must be direct-connected to the 
propeller shaft, the engine speed should not exceed from 250 to 350, as a maximum 

400, turns per minute. Between engine and propeller shaft there should be a clutch 

in order to be able to start the engine when not under load. In the case of freight 

and tug boats the clutch should be so designed as to allow of somewhat continued 

slipping of one half upon the other, in order to bring the engine up to load gradually. 
A sudden taking hold of the clutch may bring the engine up standing. Where 

reversible propellers are used this of course may be done by gradually changing the 
position of the blades. The fly-wheels for these engines should combine maximum 
moment of inertia with smallest diameter. This has often led to making them solid 
disks with wide heavy rims. It is usual to deliver marine gas engines together with 
their auxiliaries mounted on structural steel frames to the builders of the hull, to- 

1 Konstruction und Betriebsergebnisse von Fahrzeugmotoren fur flussige Brennstoffe, Berlin, 1901. 

JUhU J S F S o P r ri thf mathematical derivation of this statement see the author's treatise referred to, p. 44. 



502 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


gether with directions concerning the erection. The following extract from one of the 
older sets of “ Directions for the Erection ” of Capitaine-Grob marine engines may 

serve as an example of this practice. 

“ The engine represents a certain weight whose position must have some influence 
on the position of the hull in the water. The latter should in all cases be the natural 
one, that is, not down by the stern and least of all by the head. The factors that 
must also be considered in the problem are the effect of the cargo (passengers or 

freight) and the fact that a boat under headway always exhibits a tendency to lower 
the stern. The proper consideration of all these things determines the right location 
for the engine, which may be at the middle, that is, coinciding with the displacement 
center of gravity, or may be ahead or astern of the same. Other qualities of the 
boat, such as speed, etc., are not materially affected by engine location. 

In boats of ordinary proportions of hull, the engine is best located nearly in the 

middle. Care should of course be taken to see that the center line of the shaft comes 
directly over the keel to avoid unequal distribution of the load. 

To securely place the engine it is necessary to connect a number of frames by 
means of a foundation plate so that both engine and thrust block may be placed on 
it, thus rigidly connecting the two. 

It is of the greatest importance to have both engines and thrust block on a 

common foundation plate. 

The engine should be set as far down as the fly-wheel permits, and its position should be 
inclined enough to have the center line of the shaft coincide with that of the stem tube. 

The lower part of the fly-wheel should be encased water-tight in order to protect 
it against bilge water. 

When no special engine room or cabin is provided, that is, when the boat is 

open, it is advisable to encase the engine. To leave it without any protection whatever 
might lead to serious rusting. Where the boat carries a cabin, it pays to partition 
off a portion of it to accommodate the engine. In the larger boats of course separate 
engine rooms are provided as in steam vessels. # 

The exhaust muffler is best placed ahead of the fly-wheels between two of the 
frames. The pipe from the engine to the muffler should of course be as short as possible. 
The pipe fiom the muffler may be laid along a longitudinal wall to the stern, having near 
the end a slight incline to avoid the water. To make the exit horizontal might cause 

the gas to give trouble, neither is it advisable to exhaust under water. 

If in an open boat the motor is encased, ample opportunity should be given for 
free access to all of its parts by means of doors or covers. 

The fuel tank is in most cases best placed in the bow, and should be made as 
large as possible. In order to utilize all the available space, the tank is usually made 
to conform to the shape of the vessel and is therefore best made on the spot. The 
engine is furnished cooling water by means of the pump attached. The suction pipe 
of this pump is directly connected to a sleeve passing through the bottom of the hull. 
The warm water is discharged in a similar way. The suction opening should be 
provided with a screen to prevent the access of foreign bodies to the pump. Both 
suction pipe to pump and outlet for the jacket must be furnished with a three-way 
cock, so that the water may be shut out when the engine is to be taken apart and 
also to be able to drain the pump and jackets in winter. 

If the boat is to be used on salt water, the engine cylinder should be cooled by 
vaporization with fresh water, because the salt water is apt to deposit scale on the 
jacket walls and to corrode the metal. 


PORTABLE AND SELF-PROPELLED ENGINES 


503 


The power required by motor boats under ordinary circumstances is given in 
Table 109. The power required by tug boats in water comparatively free from current 
(canals for instance) is estimated at .1 B.H.P. for every ton of towage and for speeds 
of from 3 to 5 miles per hour. The figures, however, apply only under favorable 
average conditions; stormy weather, counter-currents, bad design of hull, etc., cause 
considerable increase in the power demand. 


Table 109 


DIMENSIONS AND PERFORMANCES OF MOTOR BOATS 


Engine 

B.H.P. 

Length of 
.Boat, Feet. 

Speed. 

Miles per Hour. 

Cargo. 

Towing 

Capacity, Tons. 

Draft, 

Ft. 

Freight, Tons. 

Passengers. 

1 

20.0 

3.8 

1.5 

8 



2 

23.0 

5.0 

2.0 

12 


1.64 

3 

26.4 

5.6 

3.0 

16 

30 

1.64 

4 

29.5 

6.2 

4.0 

22 

40 

1.97 

5 

32.8 

6.9 

5.0 

26 

50 

2.30 

8 

36.0 

8.1 

6.0 

30 

80 

2.62 

10 

39.4 

8.7 

8.0 

36 

100 

3.94 

12 

42.6 

9.3 

8.0 

36 

120 

4.10 


The fuel consumption of automobile engines varies from .16 to .24 lb. of gasoline per ton 
mile, for the average condition of road and depending upon size and quality of engine and car. 

For the larger boat engines, fuel costs alone amount to about .04 cents per 
ton mile with suction-gas engines and .20 cents with oil engines. Opposed to this 
are a cost of from .24 to .32 cents per ton mile for towing by horse or mule team 
and of from .28 to .40 cents per ton mile for electric towing. 1 

The ratio of “ steaming radius ” as between steam vessels and motor boats is from 1:6 to 


1:8(1 cu.ft. of coal is estimated to furnish 84 H.P. hours, the same volume of fuel oil supplies 
58 or 60 H.P. houis). Suction-gas installations require more space for both power plant and 
fuel than oil engines, hence the economic superiority of the suction-gas plant is the deciding 
factor. The first gas tugs were built by the Gasmotorenfabrik Deutz; lately, however, Emil 
Capitaine has also taken up this particular field of usefulness for the gas engine with a great 
deal of energy. 2 

Designs of Automobile and Launch Engines: 


Fig. 666. —Opposed Engine, Wenzel Type, 
built by Fr. Zimmermann & Co., Halle S. 

Does not differ materially from the engine 
shown in Fig. 667. Used mainly in kero¬ 
sene locomobiles, rated capacity 8-12 H.P. 



1 Note that these figures are quoted from German practice. _ 1 . 

2 See Capitaine's paper before the Schiffbautechnische Gesellschaft, 1904, printed in the year boo 

•of the Sen. G., 1905, Berlin, J. Springer. 























































































504 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Fig. 667.—Opposed (Balanced) Engine, 
Capitaine Type, built by the Leipziger 
Motorenfabrik vorm. Swiderski. 

(The combustion chamber, d, is contained 
between the heads of the opposed trunk 
pistons, which latter are connected to cranks 
180° apart. Inlet valve a is automatic, while 
the exhaust valve b is operated by an eccen¬ 
tric through two-to-one gearing, c is a throttle 
valve for air. The vaporizer acts as igniter 
and is located opposite b). This construc¬ 
tion results in good balance, so that high 
speed is permissible.* Capacity up to 12 
H.P., mostly used as a launch engine. 



Fig. 668.—Six-cylinder Loutzki Boat Engine. 

Rated H.P. = 500 at 900 r.p.m. (D = 12", 5 = 7.9"). Fuel is alcohol or 
gasoline. Weight approximately 5280 lbs. when crank case is made of “ Metorit” 
(aluminum alloy). Each cylinder has two igniters, one compressed air starting 
\alve, and one combined inlet and outlet valve, as per construction shown in 
Fig. 669. The latter valve, on account of its small lift ( = .5" while the dia.= 
11.0 ) is supposed to admit of high rotative speeds. (Valve disk a is internally 
cooled and operated from shaft a ’; concentric valve b is operated by the two 
rods b and seats at c. 



Fig. 669. —Combined In¬ 
let and Outlet Valve, 
Loutzki. 


The inventor of this construction, intended to secure balancing, is Ch. Brown, who applied it to steam 
engines (see German Patent No. 72625, 1893). Capitaine was the first to apply the principle to internal- 
combustion engines. 











































































PORTABLE AND SELF-PROPELLED ENGINES 


505 




Figs. 670-674.—Six Cylinder Loutzki Boat Engine. 


Rated Capacity 300 B.H.P. at 600 r.p.m. Weight complete 8800 lbs. Fuel consumption at 320 B.H.P. 
(max. capacity) was found to be .573 lb. of gasoline or .793 lb. of alcohol per B.H.P. Hour. 

Symbols: a, inlet valve; b, exhaust valve; c, spark plug; d, d' and e, throttle valves for air, gasoline 
vapor and mixture; /, carburetors; g, exhaust muffler, water cooled externally, internal air pre-heater, h, 
shaft governor (controls throttle valve, d '); h', commutator for ignition system, i, circulating pump, , 
vent pipe: l, oil cups for the pistons; m, magneto; n, control lever to adjust supply of cold and pre-heated 
air; o, same for the carburetors; p, governor reach rods; q, spark coils; r, interrupter, s, switc ar 






































































































































































































506 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Table 110 

DIMENSIONS AND WEIGHTS OF WENZEL AUTOMOBILE AND LAUNCH ENGINES, BUILT 
BY H. HAMPER, G. M. B. H., BERLIN, W. 

Single-cylinder Engines, Figs. 675-677 
(Dimensions in inches) 


Model. 

Rating 

B.H.P. 

R.p.m. 

Bore and 
Stroke, 
Inches. 

W’t. 

lbs. 

A 

B 

C 

D 

E 

F 

G 

H 

/ 

K 

L 

M 

N 

O 

P 

Q 

R 

A 

3 

800-1000 

3.54X3.94 

143 

8.5 

1.56 

3.9 

1.77 

18.7 

24.4 

tt 

1.97 

5.5 

4.37 

7.9 

.98 

.71 

.51 

9.3 

3.31 

5.9 

B 

4 

800-1000 

3.94X3.94 

165 

8.3 

1.56 

3.9 

1.77 

19.1 

24.8 

H 

1.97 

5.9 

4.77 

6.3 

1.10 

.71 

.50 

9.0 

3.54 

5.9 

C 

6 

800-1000 

4.33X4.73 

220 

9.6 

1.97 

4.7 

1.97 

22.4 

29.1 

li 

1.97 

6.3 

5.42 

6.8 

1.26 

.87 

.51 

10.6 

4.18 

6.7 


Table 111 


Two-cylinder Engines, Figs. 678-680 


.Mod¬ 

el. 

Rating 

B.H.P. 

R.p.m. 

W’t. 

lbs. 

A 

B 

C 

D 

E 

F 

G 

H 

/ 

K 

L 

M 

N 

O 

P 

Q 

R 

S 

T 

U 

D 

8 

750-800 

264 

7.1 

.78 

6.1 

1.77 

23. 

2 

29.3 

H 

3.9 

8.1 

12.0 

3.9 

1.42 

.98 

.51 

7.9 

1.97 

3.9 

8.3 

2.36 

15.7 

E 

10 

750-800 

319 

8.4 

.98 

5.9 

1.97 

24. 

6 

32.7 

14 

4.3 

8.9 

13.4 

4.1 

1.65 

1.18 

.51 

9.5 

1.97 

4.3 

9.1 

3.34 

15.7 

F 

12 

750-800 

440 

11.0 

1.18 

5.9 

2.36 

25. 

6 

35.4 

2 

4.3 

11.2 

GO 

CO 

5.9 

1.81 

1.18 

.63 

12.2 

2.76 

4.7 

9.8 

3.9 

19.7 


IV. Motor Locomotives and Gas Railways 

Motor locomotives (traction engines, etc.) nearly always use liquid fuel. This 
branch of the gas engine field has received bat little attention in Germany and on 
the continent in general, although many gas traction engines are in use in English 
and other colonies. It is quite likely, however, that the attempts which have for 
several years back been made by a number of the world powers to adapt the traction 
■engine to army service will do much to stimulate the further development of this 
branch of the industry. The prize competition instituted in 1901 by the Prussian 
War Office, and from which a great deal was expected, had practically no results, 
mainly because the requirements imposed upon the competitors were impossible of 









































































































PORTABLE AND SELF-PROPELLED ENGINES 


507 


fulfilment at that time. It served the purpose, however, of giving an impetus to 
the entire industry, which will very likely result in some benefit to the development 
not only of the traction engine but of the entire field of self-propelled vehicles as 

well. 

Gasoline or oil locomotives (running on rails), the building of which was to 
the author’s knowledge first taken up by G. Daimler some twenty or more years ago, 
have at the present writing reached a high state of development and are in extended 
use on industrial railways (narrow gauge) of all kinds. Their average power capacity 
does not at present exceed from 12 to 15 H.P., except in rare cases. But a good 
solution of the problem of direct reversibility would also do a great deal toward 
further development in this case. 

Gas locomotives burn illuminating gas compressed to from 120 to 225 lbs. per 
sq.in. This gas is carried in tanks, which hold sufficient fuel for certain limited 
distances, the pressure being reduced before use to nearly atmosphere by means of 
automatic regulators. 


Fig. 681. —Main Dimensions and Capacities of Motor 
Locomotives, Fig. 681, built by Capitaine-Grob Co. 
Normal speed 5 miles per hour. 

The weight in service depends upon the rail gauge. 
A change in gauge changes the weight. 

The total loads moved (exclusive of weight of engine), 
as given below, are based upon the assumption that the 
Cars as well as the track are in fair condition, and that 
the weather conditions are normal. 



Table 112 



4 

6 

8 

10 


2 

2 

2 

2 

Number of driving axles. 

Diameter of wheels, ft. 

1.97 

1.97 

2.30 

2.30 


33.5 

33.5 

35.4 

37.5 

wneei uase, .. 

17.7 

19.7 

19.7 

23.6 

Minimum gauge, inches. 

110 

110 

165 

220 

ouppiy Ol UUUlilig vvaici, . 

nllnns. 

5.2 

6.5 

9.1 

10.4 

•» r l _*■**-» + ton limirs) (T 

ouppiy oi Kerosene (sumbicni 0 - 

•*rr • i . 1 i! . 1 TlOl 1 ( 1 <3 . 

3520 

4400 

5500 

7270 

Weignt Ol locomotive m 


275 

374 

517 

725 

inflective tractive enon, pouiiua. 

[1:20 =50% 

.8 

1.0 

1.5 

2.0 


1:25 =40 

1.0 

1.5 

2.0 

3.0 


1:30 =33.3 

1.5 

2.0 

3.0 

4.0 


1:40 =25 

2.0 

3.0 

4.5 

6.0 

Total loads moved in tons (exclusive of. 

1:50 =20 

3.0 

4.5 

7.0 

8.0 

1:60 =16.5 

4.5 

6.0 

8.0 

10.0 

weight of locomotive) on grades of 

1:80 =12.5 

6.0 

7.5 

10.0 

13.0 

1:100 = 10 

7.0 

9.0 

13.0 

16.0 


1:200 = 5 

11.0 

14.5 

20.0 

25.0 


1:500 = 2 

15.0 

21.0 

29.0 

35.0 


. 1: oc =0 

20.0 

28.0 

39.0 

50.0 


























































508 CONSTRUCTION, ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 



Fig. 082.—Deutz Motor Locomotive. 

The smaller sizes of Deutz gasoline 
mine locomotives differ somewhat re¬ 
garding the construction of the trans¬ 
mission gear, etc., from the types 
shown in Figs. 682-684. The external 
dimensions are such that the com¬ 
pletely erected locomotive can pass 
through the mine shaft. (See Z. d. 
V. D. I., 1902, p. 490). 


Table 113 

PRINCIPAL DIMENSIONS AND CAPACITIES OF DEUTZ LOCOMOTIVES 


Type 


Size of engine, B.H.P. 

Tractive effort on level 
track, based on a 
speed of m. per hr.. . 

= appr. lbs. 

Length inch buffer, ft.. 

Minimum width, ft. . . 

Height without roof, ft. 

Minimum gauge, in. . . 

Weight in service, sin¬ 
gle transmission, lbs. 

Do., double transmis¬ 
sion, lbs. 

* With roof. 


Mine or 
Agricul¬ 
tural 
Railway. 

Mine or 
Agricul¬ 
tural 
Raiiway. 

Mine or 
Agricul¬ 
tural 
Railway. 

Street 
Railway 
or Switch 
Engine. 

Mine or 
Agricul¬ 
tural 
Railway. 

Street 
Railway 
or Switch. 
Engine. 

Mine or 
Agricul¬ 
tural 
Railway. 

Street 
Railway 
or Switch 
Engine. 

Street 
Railway 
or Switch 
Engine. 

6 

8 

12 

16 

24 

32 

2.5-4.5 

2.5-5.6 

2.5-7.5 

2.5-7.5 

2.5-7.5 

2.5-7.5 

2.5-7.5 

2.5-7.5 

2 .5 / . 5 

572-330 

770-352 

1152-384 

1100-352 

1560-505 

1540-462 

2420-770 

2420-770 

3180-1010 

9.05 

9.80 

11.65 

10.5- 

11 .80 


13.10 

11.80- 

13.10- 




12.30 




13.10 

14.45 

2.63 

2.86 

3.35 

6.65 

3.78 


4.43 

7.23 

7.23 

5.02 

5.08 

5.42 

8.85* 

5.91 


6.43 

8.85* 

8.85 * 

17.8 

18.7 

19.7 

29.5 

23.6 

29.5 

23.6 

29.5 

29.5 

5940 

6800 

9700 








8150 

10 500 

14 200 

12 300 

16 500 

15 400 

19 800 

25 300 


The first gas street railway was put in operation in Dessau in 1894. The cars used 
were furnished with a two-cylinder opposed (Luhrig) engine which was placed under one 
of the seats lengthwise of the car so that the fly-wheel came about the middle and 
outside of the car frame. 1 The power originally used was 8 H.P. which was afterward 
increased to 12 H.P. This was transmitted to the driving axles at various speeds 
by means of a compound gear transmission. The gas tank was located under the 
opposite seat, while the cooling water tank was placed on the roof. The gas supply 
carried amounted to about 28 cu.ft. under a pressure of about 120 lbs., when this 
was nearly used up the cars had to stop at special compressing stations for a new 
charge. The motive power on this road was changed to electricity a few years ago. 


1 Drawings and description in Z. d. V. D. I., 1895, p. 1009. 




































PORTABLE AND SELF-PROPELLED ENGINES 


509 


Operating Results, (a) The gas consumption on the Dessau Street Railway 

amounted to from 11.5 to 17.7 cu.ft. of gas per car mile, varying with the grades 

on the particular run. The average speed was 6.25 miles per hour including stops. 

(6) According to data furnished by Kemper, 1 the cost of building gas railways is 
about $30,000 per mile, and the total operating costs about 6.5 cents per car mile 
(assuming the price of gas at 85 cents per 1000 cu.ft.) Compared with this, the 
same authority states the cost of building an electric road to be about $36,000 per 
mile, and places the total operating costs at 7.8 cents per car mile. In spite of this 
apparently superior showing, gas railways have found very little application. The 
reasons for this probably have nothing to do with the economic side of the case, 
but are more likely found in certain external difficulties and inconveniences experienced 
in operation. The gas engine makes larger demands as far as space is concerned 
than the electric motor, and further, although it may be comparatively slow running, a 
reciprocating engine causes vibrations still distinctly felt in spite of nearly perfect 
balancing. The heat radiated lay it, as well as gas and lubricating oil vapors, are 
apt to cause annoyance in the interior of the car, while the exhaust gas and oil 

thrown around may do the same thing on the outside. This was particularly noticed 
in the case of the Dessau railway by persons who had been used to electric cars. 

(c) A self-propelled car often found on small private railways is the Daimler 

gasoline passenger car. This car, being independent of any charging station, is therefore 
more simple to operate and shows better results than the car operated with gas. 
This is quite clearly shown by the following table, which is taken from a report on 
Daimler cars made by the Kirchheimer Eisenbahngesellschaft: 


Table 114 

GASOLINE COSTS, DAIMLER NARROW-GAUGE GASOLINE CAR 


Weight of car with 5 to 6 B.H.P. engine. . 7050 lbs. 

Wheel base. 4 .60 ft. 

•Contains seats for. 10 

and standing room for. 10 

"Trial trip between Esslingen and Plochin- 
gen, with a load of 7 passengers and 1340 
lbs. of iron, equivalent to 15 passengers. 

Length of line. 5.62 m., 


of which 4.37 miles are level, while the remaining 
1.25 miles show grades of ^5 to 


Trip out: Time 37 min. corresponds 


to a speed of approximately.9.4 miles per hr. 

Gasoline used: In vaporizer.416 gallon 

In ignition lamp.052 “ 


Total.468 gallon 

Hence cost per mile= —— X 20 = 1.62 cents. 

D.D 2 * 

Return trip: Time 34 min., equivalent 

to a speed of, approximately. 10 miles per hr. 

Gasoline used: In vaporizer.350 gallon 

In ignition lamp.052 “ 


Total.402 


Hence cost per mile = —— X20 — 1.43 cents. 
5.62 


(d) The computation of operating costs presented by the following table is given 
by the Gasmotorenfabrik Deutz for the type of locomotives illustrated by Figs. 
683 and 684. They are based upon the actual expenses connected with the operation 
of two mine locomotives of 6 and 8 H.P. respectively, and of two farm railway 
locomotives of 8 and 12 H.P. 


1 Journal fur Gas Beleuchtung, 1893, p. 505. Note that these and the following figures apply to 
■German practice. 





















510 CONSTRUCTION ERECTION, TESTS OF INTERNAL-COMBUSTION ENGINES 


Table 115 

OPERATING COST OF DEUTZ GAS-ENGINE LOCOMOTIVES 


Owner. 

Graflich v. 
Arminsches 
Forstamt. 

Vereinigte Konigs- 
und I.aurahutte, 
Akt. Ges. 
Laurahutte. 

Verein. Gesellschaft 
fur Steinkohlenban 
in Wurmrevir, 
Kohlscheid. 

Herzogl. Wiirttem- 
bergisches Rentamt 
Karlsruhe, O.-S. 

Size . 

6 H.P. 

8 H.P. 

8 H.P. 

12 H.P. 

How used. 

Mine 

Mine 

Farm railway 

Brickyard railway 

Length of haul, miles. 

Grade . 

Load, tons . 

Capacity, ton miles. 

.375 

Level 

93^8 } 10 hr. shift 

.28 

Level 

} 9 ? hr. shift 

.485 

1:62.5 

Jqq | 10 hr. shift 

1.87 

1:33 

ogQ | 11 hr. shift 

First cost: 

Locomotive, dollars. 

Charging stations, dollars. 

Type Cl. 

1675 

75 

Type CI. 

1875 

75 

Type CII. 

2135 

112.5 

Type CII. 

2490 

512 

1750 

1950 

2247.5 

3002 

Yearly costs: (300 days) 

Interest, depreciation, repair, 

17%, dollars. 

Engineer, dollars. 

Fuel, dollars. 

Oil and waste, dollars. 

298 

300* 

135 

42.5 

331 

300 

202 

46 

382 

300 

190 

45 

510 

300 

256 

53.5 

775.5 

879 

917 

1119.5 

Total operating costs per ton mile, 
cents . 

2.77 

2.62 

2.81t 

1.33 


* Very low for American conditions. 

f The comparatively high cost is directly due to small demand on the locomotive. 


Compared with the above table, the cost of doing the same work with horses would 
have been from 4 to 8 cents per ton mile, depending upon locality and hauling distance. 
Designs of Gasoline Locomotives: 













































































































































































THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


PART IV 


A. FUELS 


In the operation of internal-combustion engines we consider only the fuels that will 
form ready mixtures with air and which will burn without any considerable residue. 
These requirements are most nearly met by the gas fuels, and for that reason these 

I have been almost exclusively used in the operation of engines even from the first- 
stages of the development. 

Natural gas, which is obtained from wells in various localities of the world, need 
not be considered here. 1 Most of our fuel gases are obtained either through distillation 
or through gasification of solid fuels. The fuel oils are converted into vapor or spray 
before being burned in engines. Between the original condition of the fuel and the 
state in which it is finally used there is consequently an act of transformation which 
may be brought about in apparatus either entirely independent of the engine instal¬ 
lation, in apparatus clearly auxiliary to the engine, or even in the engine itself. 
Examples of these three methods of transformation are found in the manufacture 
of illuminating gas, the making of producer gas, and the production of kerosene 
vapor. If the transformation takes place outside of the engine, the process usually 
involves, through loss of fuel and heat and through operating cost, an increase in 
the specific cost of heat of the fuel consumed. Any fuel consequently possesses its 
lowest heat cost in its original shape, and other things being the same, would therefore 
give the highest economy when used directly in this condition. In the operation of 
gas engines, however, this is possible only with liquid fuels. 

The choice of solid fuels suitable for the production of power gas is, generally 



1 Translator’s Note. This gas is yet largely used in 
are given on page 526. 


used in American practice, and hence some constants 


511 






512 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


In this respect the fuels consisting most nearly of pure carbon, that is, anthracite 
and coke, maintain their position as the most common fuels used in producer practice. 
All of the bituminous coals, when used in producers, as yet cause serious difficulties, 
either through high percentage of ash, presence of sulphur and tar-forming hydro¬ 
carbons, tendency to coke or clinker, irregular size, etc. However, the building of 
serviceable producers to handle the so-called brown coals or lignites, especially in the 
shape of briquettes, has been quite successful (see p. 282). 

The following tables show the composition of a number of different fuels together 
with their heating values. The first table, No. 116, has been compiled entirely from 

Table 117 


Kind of Fuel. 

Name or Location. 

Proximate Analysis. 

Heating Value per lb. 

Moisture, 

% 

Vol. Matter, 

% 

Fixed C, 

% 

Ash. 

% 

of 

Combustible 

of 

Dry Coal 

Pa. anthracite 

Treverton. 

.84 

6.67 

85.66 

6.83 

15195 

14149 

£ £ 

Lykens Valley (buckwheat) 


6.80 

80.20 

13.00 

13680 

11901 

£ £ 

Jermyn (stove). 


6.08 

82.90 

11.02 

13850 

12324 

£ £ 

Lackawanna. 


5.00 

84.00 

11.00 

13900 

12371 

f £ 

Avondale . 


6.00 

87.78 

6.19 

14200 

13317 

£ £ 

Manville Shaft, Scranton. .. 


6.12 

86.50 

7.38 

14100 

13059 

Pa. bitnminmis 

Antium. 


18.54 

70.16 

11.30 

15400 

13660 

i £ 

Beaver Creek. 

1.50 

34.33 

55.42 

8.75 

14762 

13450 

tt 

Big Muddy. 


83.20 

58.20 

8.60 

14700 

13436 

i ( 

Carnegie. 

1.45 

36.42 

56.20 

5.93 

14947 

14047 

it 

Creedmoor. 

1.09 

38.91 

51.14 

8.86 

14983 

13641 

(i 

Pittsburg. 


35.50 

54.60 

9.90 

14200 

12794 

i < 

Reynoldsville. 


24.67 

69.96 

5.37 

16100 

15235 

t( 

Youghiogheny. 

1.40 

33.13 

60.82 

4.65 

16031 

15275 

W. Va. bit. 

Clover Hill. 

1.34 

31.70 

56.83 

10.13 

14265 

12800 

£ £ 

Pocahontas . 

.80 

18.30 

73.65 

7.25 

15682 

14536 

Ohio bitumin’s 

Cambridge. 

2.43 

37.79 

50.36 

9.42 

14474 

13076 

( £ 

East Palestine. 

.82 

34.98 

52.65 

11.89 

14603 

12902 

£ £ 

Hocking Valley. 

6.65 

34.14 

49.54 

9.67 

13972 

12524 

< f 

Mahoning. 

3.15 

35.00 

50.95 

10.90 

14578 

12416 

£ £ 

Waterford. 

1 .55 

37.29 

53.34 

7.82 

14814 

13637 

Ind. bitumin’s 

Brazil. 

8.98 

34.49 

50.30 

6.28 

14542 

13546 

£ £ 

Lancaster. 

12.66 

37.44 

47.22 

2.68 

14251 

13813 

£ £ 

New Pittsburg. 

5.89 

42.23 

40.40 

11.48 

13973 

12269 

Illinois biturn’s 

Bryant . 

2.42 

32.91 

42.64 

22.03 

13157 

10187 

£ £ 

Big Muddy. 

7.40 

28.30 

53.90 

10.50 

13757 

12212 

( £ 

Gillespie. 

12.60 

30.60 

45.30 

11 .50 

12510 

11947 

£ £ 

Ledd. 

8.49 

33.53 

44.12 

13.76 

13876 

11769 

£ £ 

Mt. Olive. 

8.00 

36.40 

44.50 

11.10 

13158 

11570 

( £ 

Peru. 

9.00 

37.20 

47.20 

6.6 

13603 

12616 

£ £ 

Streator . 

12.10 

35.30 

48.80 

3.90 

13580 

12993 

£ £ 

Vulcan. 

10.30 

27.90 

50.00 

12.80 

13817 

11999 

( ( 

Wilmington (screenings) .. . 


32.60 

39.70 

27.70 

13200 

9544 

Lignite. 

Erie, Colorado . . . 

18 57 

32 71 

45 98 

2 74 

11360 

10978 

£ £ 

Canon City, Colorado. 

6.56 

36.74 

47.93 

8.76 

11170 

10122 

£ £ 

Marshall, Colorado. 

13.19 

37.84 

46.43 

2.54 

11478 

11143 

Oven coke.... 

Connell sville, Pa. 

0.03 

0.46 

89.58 

9.11 

14211 

12801 

£ £ 

Fairmont, Pa. 

0.30 

0.62 

85.78 

11.46 

14464 

12534 

£ £ 

Irivin, Pa. 


1 .38 

88 24 

9 41 

14447 

1°947 

£ £ 

Kittanning, Pa. 



87 22 

11 43 

14483 

1263 9 

£ £ 

New River, W. Va. 



92.38 

7 21 

14427 

I 3395 

( £ 

Pocahontas, W. Va. 

0.29 

0.49 

92.58 

6.05 

14400 

13441 

£ £ 

Birmingham, Ala. 



87.29 

10.54 

14290 

12474 















































































Table 1 


State. 

County. 

Locality. 

Alabama 

Walker. 

Horse Creek ... 

l ( 

i i 

Carbon Hill .. 

Arkansas 

Sebastian 

Huntington. 

i t 

11 

1 Bonanza . 

t ( 

11 

j Jenny land. 

1t 

Colorado 

Johnson ... 
Boulder. 

Coal Hill 

Lafayette . . . 

Illinois 



< l 


t i 

t ( 

« ( 

tt 

Williamson 
Madison. 

t t 

Marion . 

Troy . 

Collinsville 

( t 

Montgomery . . . 

Coffeen. 

Indiana 

Sullivan. 

Mildred . . . 

l t 

Warrick. 

Boonville . 

hid. Terr. 


Hpnryptta 

i t 


Hartshornp 

i t 


Ed wa rd s 

( l 


Lehigh . 

t t 


11 

t t 


Coalga tp . 

Iowa 

Wapello. 

Laddsdale . 

i t 

Marion .... 

Liberty Township . . . 
Altoona . 

tt 

Polk. 

tl 

Appanoose. 

Centerville. 

i l 

Lucas. 

Charlton . . 

Kansas 

Crawford. 

Fleming . . 

“ 

i t 

Yale .... 

t t 

Cherokee. 

Shannon. . . 

t ( 

Atchison. 

Atchison . 

( t 

Cherokee . 

West Mineral . 

Kentucky 

t i 

Bell. 

Straight, Crppk 

Hopkins. 

Earlington. . 

i C 

i i 

Barnsley . 

tt 

Webster. 

Wheatcroft . . 

Missouri 

i t 

Bates. 

Macon. 

New Home. 

Bevier . . 

tt 

Putnam. 

Mendota .. . 

11 

Morgan. 

Versailles .. 

Montana 


Red Lodgp.. 

New Mex. 

McKinley. 

Gallup .. . 

11 

11 

t t 

N. Dakota 

Stark. 

Lehigh. 

11 


Wilhston. 

Pa. 

Somerset. 

Windber. 


Owner or Operator. 


Ivy Coal & Iron Co. . . 

Galloway Coal Co. 

Central Coal & Coke Co. 


Western Coal & Mining Co. 


Coke Co. . 


i C 
tc 
u 
u 

It 

it 

li 

tt 


Texas 


W. \ 


a. 


Washington . . 


Scranton . 

Meadow Lands 

Anderson. 

Hackett. 


Ellsworth 


Manifold. 

Allegheny. Murdocksville... 

Washington . . . . j Midland. 

Wotters Station. 

Wood. Hoyt. 

Marion.j Ivingmont. 

Clarksburg. 


Preston . . 
Randolph 
Fayette . . 


Wyoming 


McDowell 

l t 

Sheridan . 


Morgantown 

Bretz . 

Coalton .... 
Rush Run . . 

Sun . 

Ansted . 

Powellton . . 

Zenith. 

Big Sandy . . 
Sheridan . . . 
Cambria . 


Southern Ill. Coal Mining & Washing Co. 
Donk Bros. Coal & Coke Co. 


Clover Leaf Coal Co. 

J. Wooley Coal Co. 

T. D. Scales Coal Co. 

Whitehead Coal Mining Co. 

Rock Island Coal Co. 

D Edwards & Son. 

Western Coal & Mining Co. 


Southwestern Development Co. 

Anshoe Coal Co. 

Mammouth Vein Coal Co. 

Gibson Coal Mining Co. 

Centerville Block Coal Co. 

Inland Fuel Co. 

Western Coal & Mining Co ... . 


Southern Coal & Mercantile Co 

Atchison Coal Mining Co. 

Southwestern Development Co. 
National Coal and Iron Co ... . 
St. Bernard Mining Co. 


Kind of Coal. 


Mine Number or Name. 


Bit., lump and nut. 

Bit., lump, nut, and pea. . 

Bit., lump and nut. 

Bit., lump. 

Bit., lump and slack .... 
Bit., lump and slack .... 
Run of mine, black lignite 
Bit., lump and slack .... 

Bit., slack. 

Bit., run of mine. 

Bit.., lump. 

Bit., washed slack. 

Bit., run of mine. 

Bit., run of mine. 

Bit., run of mine. 

Bit., lump and slack .... 

Bit., run of mine. 

Bit., run of mine. 

Bit., lump. 

Bit.., slack and pea. 

Bit., slack. 

Bit., lump and fine coal . . 

Bit., run of mine. 

Bit., lump. 

Bit., lump. 

Bit., run of mine. 


8 

Chickasaw, No. 5 . . 

3 

12 

18 

4 

Simpson. 

Nigger Hollow. 


3 

3 

1 

Clover Leaf, Sft. No. 1 

Mildred. 

Electric. 

1 

8 

1 

r. 


o 

4 

3 

Inland No. 1 
10 


Wheatcroft Coal & Mining Co. . . 

New Home Coal Co. 

Northwestern Coal & Mining Co. 

Mendota Coal & Mining Co. 

Morgan County Coal Co. 

Northwestern Imp. Co. 

American Fuel Co. 

Caledonian Coal Co. 

Consolidated Coal Co. 

Cedar Coules Coal Co. 

Berwind White Coal Mining Co. 

it t t 

Pennsylvania Coal Co. 

McLain. 

Pittsburg-Buffalo Co. 


Bit., lump, nut, and slack . .. 

Bit., run of mine. 

B>t.., lump. J Atchison 

Bit., lump and nut.! 

Bit.., run of mine. 

Bit., lump, nut, pea, and slack 

Bit., run of mine.. . 

Bit., run of mine. 

Bit., run of mine. 

Bit.., run of mine. 

Bit., slack. 

Bit., run of mine. 

Black lignite, washed nut. 

Bit., lump and slack. 

Bit., slack. 

Run of mine, brown lignite . . . 

Brown lignite. 

Bit., run of mine. 

Bit., run of mine. 

Anthracite culm. 


11 

9 


11 

2 

11 

Barnsley. .. 
Wheatcroft . 
New Home . 

8 

Mendota . . . 


Weaver. 

Otero. 

Lehigh. 

Cedar Coulee.. 
Eureka No. 31. 

t t tt 


Pittsburg Coal Co. 

James Ellsworth & Co . . . 
James W. Ellsworth & Co. 


Pittsburgh Coal Co. . . 

Houston Co. Coal <fc M’f’g Co. 

Consumers’ Lignite Co. 

Va. and Pittsburg Coal & Coke Co. 

Pitcairn Coal Co. 

West Virginia Coal Co. 


Davis Colliery Co . 

New River Smokeless Coal Co. 


Ganley Mountain Coal Co. . 

Mt. Carbon Coal Co . 

W. H. Coffman . 

Big Sandy Coal & Coke Co. 
Wyoming Coal Mining Co. . 
Cambria Fuel Co . 


Brown lignite 
Brown lignite 
Bit., 

Bit., 

Bit., 

Bit. 

Bit.. 

Bit., 

Bit., 

Bit., 

Bit., 

Bit., 

Bit., 


McLain’s. 

Blanche . 

Russell. 

Nottingham . . . . 
Colliery No. 1. . . 
Colliery No. 2. . . 

Manifold. 

Natural Outcrop 
Mine No. 3. 


Page 

of 

Report. 


I 


1 and 2 

run of mine. | Kingmont. . . 

run of mine.Pitcairn . 

run of mine.| Richard. 

run of mine.Bretz. 

lump and nut.! Coalton 

run of mine. 

run of mine . 

run of mine. 

run of mine. 

run of mine. 

run of mine. 

Black lignite. 

Bit., run of mine . 


Rush Run. 

Sun No. 1. 

Ganley Mountain . .. 

Vulcan .. 

Zenith, 1 and 2. 

Big Sandy. 

Monarch. 

Antelope and Jumbo 


190 

197 

198 

199 

200 
202 
204 
20G 

207 

208 

209 

210 
211 
212 

213 

214 

215 

217 

218 

219 

220 
221 
222 

223 

224 

225 

226 

227 

228 

229 

230 

231 

232 

233 

234 

235 

236 

237 

238 

239 

240 

241 

242 

243 

244 

244 

245 
272 
272 
272 
272 
272 

272 

273 
273 
273 

246 

247 

248 

249 

250 

251 

252 

254 

255 

256 

257 

259 

260 
262 
263 


Coal, Car Sample. 


Coke, Sample from Washed Coal. 


Proximate Analysis. Per Cent. 

Ultimate Analysis, Per Cent. 

B.T.U. 

per 

Pound 

1 a? 

Received. 

Proximate Analysis, Per Cent. 

Sulphur 
in Coke, 
Per Cent 

Moisture. 

Vol. 

Matter. 

Fixed 

C. 

Ash. 

H 

C 

N 

O 

S 

A*h. 

Moisture. 

Vol. 

Matter. 

Fixed. 

C. 

Ash. 

2.34 

31.84 

53.28 

12.54 

5.01 

71.58 

1.65 

8..50 

.72 

12..54 

12856 






3.36 

32.88 

51.33 

12.43 

4.84 

68.69 

1.54 

11.49 

1.01 

12.43 

12350 






3.24 

17.46 

66.69 

12.61 

4.15 

74.09 

1.44 

6.47 

1.24 

12.61 







2.23 

16.02 

72.55 

9.20 

4.24 

73.83 

1.38 

4.48 

1.87 

9.20 

13750 






2.19 

19.47 

66.71 

11.63 

4.17 

75.31 

1.53 

6.08 

1.28 

11.63 

13464 

1.05 

2.80 

72.73 

23.42 

1.52 

2.36 

12.68 

72.88 

12.08 

3.82 

76.44 

1.37 

4.30 

1.99 

12.08 

13259 






18.68 

34.88 

40.45 

5.99 

6.07 

57.46 

1.15 

28.78 

.55 

5.99 

10143 





.... 

9.75 

37.48 

39.57 

13.20 

5.31 

59.72 

1.03 

16.64 

4.10 

13.20 

11025 





.... 

12.03 

31.86 

33.67 

22.44 

5.04 

50.22 

.72 

17.58 

4.00 

22.44 

9149 

1.57 

2.83 

75.42 

20. 18 

2.75 

8.50 

29.47 

50.75 

11.28 

5.09 

65.48 

1.39 

15.04 

1.72 

11.28 

11776 

6.11 

.42 

82.25 

10.92 

1.13 

12.91 

31.90 

43.55 

11.64 

5.43 

60.74 

1 15 

19 72 

1.32 

11.64 

10804 


.... 



.... 

17.02 

30.60 

35.59 

16.79 

5.50 

50.77 



3.29 

16.79 

9319 


.... 



.... 

14.43 

29.48 

42.81 

13.28 

5.49 

54.59 

1.11 

21.52 

4.01 

13.28 

10064 



.... 


.... 

11.40 

33.81 

41.39 

13.40 

5.37 

60.34 

1.18 

17.21 

2.50 

13.40 

11061 

5.71 

1.18 

80.52 

12.59 

1.69 

9.62 

36.14 

41.22 

13.02 

5.33 

60.70 

1.20 

15.32 

4.43 

13.02 

11122 






7.04 

34.55 

48.40 

10.01 

5.34 

67.55 

1.25 

13.93 

1.92 

10.01 

12202 


.... 




4.45 

36.15 

48.40 

11.00 

5.17 

69.49 

1.67 

11.15 

1.52 

11.00 

12607 

0.96 

2.59 

85.33 

11 12 

1.75 

4.61 

37.00 

47.25 

11.14 

4.92 

67.37 

1.48 

11.46 

3.63 

11.14 

12319 


.... 




6.24 

37.26 

43.29 

13.21 

4.93 

62.34 

1.36 

14.20 

3.96 

13.21 

11228 






8.29 

30.61 

36.05 

25.05 

4.37 

50.98 

1.19 

14.46 

3.95 

25.05 

9110 






8.03 

31.28 

41.40 

19.29 





3.20 



• • • • 





8.24 

30.74 

45.02 

16.00 

4.81 

59.82 

.94 

13.40 

5.03 

16.00 

11027 

10.53 

1.63 

70.39 

17.45 

3.89 

14.21 

33.17 

37.40 

15.22 

5.50 

54.08 

1.31 

19.23 

4.66 

15.22 

10019 



.... 



13.88 

36.94 

35.17 

14.01 

5.52 

54.68 

.84 

18.80 

6.15 

14.01 

10244 

5.73 

1.87 

75.49 

16.91 

4.57 

14.08 

35.59 

39.37 

10.96 

5.57 

58.49 

.90 

19.82 

4.26 

10.96 

10723 

13.05 

2.32 

73.10 

11.53 

2 97 

15.39 

30.49 

41.49 

12.63 

5.74 

55.81 

1.14 

21.49 

3.19 

12.63 

10242 






4.99 

32.68 

49.36 

12 97 

4.98 

67.34 

1.08 

9.35 

4.28 

12.97 

12242 






4.18 

31.23 

46.68 

17.97 

4.69 

61.88 

.92 

8.33 

6.27 

17.91 

11642 






2.50 

33.80 

51.25 

12.45 

4.91 

69.07 

1.20 

6.69 

5.68 

12.45 

12900 



.... 



6.95 

35.70 

45.16 

12.19 

5.25 

62.74 

1.04 

10.74 

8.04 

12.19 

11905 

.52 

1.68 

79.82 

17.98 

6.15 

4.10 

31.65 

53.71 

10.54 

5.10 

70.25 

1.06 

9.28 

3.77 

10.54 i 

12895 


.... 




3.10 

36.12 

56.39 

4.39 

5.43 

77.37 

1.83 

9.76 

1.22 

4.39 

14148 

1.50 

.83 

92.32 

5.35 

.86 

7.91 

37.94 

45.02 

9.13 

5.48 

65.81 

1.22 

14.74 

3.62 

9.13 

12200 





.... 

7.92 

36.09 

45.93 

10.06 

5.39 

65.29 

1.40 

14.34 

3.52 

10.06 

12022 

.14 

.56 

86.31 

12.99 

2.16 

5.27 

35.07 

45.48 

14.18 

4.71 

64.65 

1.24 

10.68 

4.54 

14.18 

11950 

• • • • 





8.33 

23.58 

38.73 

19.36 

4.97 

57.00 

.94 

12.48 

5.25 

19.36 

10586 






11.50 

33.63 

38.01 

16.86 

5.12 

54.79 

.96 

17.11 

5.16 

16.86 

10179 

3.45 

1.80 

80.27 

11.48 

2.79 

15.71 

28.62 

34.89 

20.78 

5.23 

48.87 

.82 

20.61 

3.69 

20.78 

8840 






12.67 

41.45 

41.05 

4.83 

6.18 

66.87 

.69 

16.31 

5.12 

4.83 

12487 

2.51 

1.11 

I- 

1C 

lO 

00 

10.81 

4.60 

11.05 

35.90 

42.08 

10.97 

5.37 

59.08 

1.33 

21.52 

1.73 

10.97 

10539 






12.29 

34.58 

46.14 

6.99 

5.82 

63.31 

1.03 

32.22 

.63 

6.99 

11252 






10.79 

33.82 

36.73 

18.66 

5.22 

55.07 

.95 

18.84 

1.26 

18.60 

9907 






35.38 

29.59 

25.68 

9.35 

6.61 

40.23 

. 54 

41.72 

1.55 

9.35 

6923 






36.78 

28.16 

29.97 

5.09 

6.93 

41.87 

.69 

44.94 

.48 

5.09 

7204 






1.10 

15.80 

75.69 

7.41 

4.20 

81.98 

1.36 

3.56 

1.49 

7.41 

14499 






0.59 

16.61 

76.76 

8.04 

4.28 

83.94 

1.27 

3.56 

.91 

6.04 

14753 






5.41 

7.02 

71.79 

15.78 

3.10 

72.65 

. ( 1 

6.96 

.74 

15.78 

12047 






1.90 

36.20 

53.70 

8.20 





1.52 








1.70 

37.20 

55.83 

5.27 




.... 

1.13 


14335 






1.46 

35.56 

53.39 

9.59 





2.05 








1.72 

36.98 

56.55 

4.75 





1.15 








1.22 

36.28 

56.24 

6.26 





•84 


14247 






1.05 

36.65 

57.25 

5.05 





.91 


.... 






1.37 

37.10 

53.84 

7.69 

.... 




1.61 








2.15 

39.15 

52.65 

6.05 

, , , , 


.... 


3.64 








3.63 

34.23 

51.12 

11.02 



.... 


1.90 



.... 





34.70 

32.23 

21.87 

11.20 

6.93 

39.25 

.72 

41.11 

.79 

11.20 

7056 






33.98 

31.01 

27.33 

7.68 

.... 




.56 








1.75 

36.77 

55.14 

6.34 

5.28 

78.00 

1.54 

7.94 

.90 

6 34 

14107 






1.95 

39.94 

50.25 

7.86 

5.13 

74.07 

1.36 

8.10 

3.18 

7 . 86 

13790 






2.29 

29.86 

57.62 

10.23 

4.99 

75.13 

1.42 

7.17 

1.06 

10.23 

13588 






1.48 

28.58 

61.55 

8.39 

4.89 

77.82 

1.48 

6.52 

.90 

8.39 

14069 






1.45 

28.97 

59.48 

10.10 

4.83 

75.75 

1.47 

6.87 

.98 

10.10 

13718 






1.53 

21.54 

71.88 

5.05 

4.76 

82.87 

1.68 

4.99 

.65 

5.05 

14807 






3.94 

19.88 

71.25 

4.93 

4.60 

79.78 

1.01 

8.52 

1.16 

4.93 

14382 






4.16 

31.28 

57.39 

7.17 

5.32 

76.70 

1.34 

8.57 

.90 

7.17 

13786 






4.08 

28.61 

60.73 

6.58 

5.23 

76.89 

1.58 

■8.95 

. 77 

6.58 

13925 






4.07 

16.34 

68.47 

11.12 

4.27 

76.51 

1.00 

6.59 

.51 

11.12 

13509 






1.72 

17.85 

73.56 

6.87 

4.43 

82.71 

1.33 

3.98 

.68 

6 . 87 

14751 






22.63 

35.68 

37.19 

4.50 

6.39 

54.91 

1.02 

32.59 

.59 

4.50 

9734 






9.41 

35.02 

34.82 

20.72 

5.00 

51.46 

.74 

18.17 

3.91 

20.72 

9650 



.... 



































































































































































































































































































































































































































































. 












FUELS 


513 


11 Professional Paper No. 48, Report on Operation of the Coal Testing Plant of the 
U. S. Geological Survey at St. Louis, 1904.” 

Table 117 contains some additional information on Pennsylvania, Ohio, Indiana, 
and Illinois coals, together with a few figures on coke, lignites, and peat. This table 
has been compiled from Poole, “ The Calorific Power of Fuels.” The reader is referred 
to this book, which contains extensive tables of the properties of all kinds of fuel. 

In the coking process some of the caibon in the original coal is always lost, as 
a part of it is used to form the volatile gases in conjunction with hydrogen (tar 
and hydrocarbons in the illuminating gas). The distillation commences at from 900 
to 1450° F., and is therefore probably completed before the coal has reached the 
gasification zone of the producer, at least with small sized fuel. The main constituent, 
therefore, that is concerned in the combustion part of the gasification process is the 
carbon in the resulting coke, and this is often, as for instance in the case of lignite 
and peat, but a small percentage of the original fuel. 

Consequently in such cases the yield of gas per pound of fuel is correspondingly low. 
High moisture content in the two fuels mentioned also sets a limit to their successful use 
in generators, see p. 269, and Section 5 in the Appendix. 

Where fuel must be stored in large quantities or transported over considerable, 
distances, in certain cases by the motor itself, the question of heat density, that is 
the heat units contained in unit volume of the fuel, assumes importance. 

Heat density is, roughly, inversely proportional to relative volume occupied, from 
which it follows that in this respect the ' liquid fuels stand first and the .gas fuels 
last. It is of course possible to increase the heat density of gases by compression, 
but that by no means establishes anything like equality between them and the fuel 
oils. According to Table 118, for instance, rich oil gas would have to be compressed 
to 14 250 lbs. per sq.in., and illuminating gas even to 25 000 lbs. per §q.in., to make 
them volume for volume the equal of kerosene. From this it follows that where fuels 
must be stored or transported the liquid fuel is the ideal. 


Table 118 

WEIGHT, HEAT DENSITY, AND HEAT COST OF VARIOUS FUELS 


VV JDAVJ1-L JL , ^ ~ - 


Gas Fuels. 

Liquid Fuels. 

Solid Fuels. 


Illuminating 

Gas. 

a 

o 

o 

Producer Gas. 

Blast-furnace 

Gas. 

6 

c 

da 

2 

s 

M 

© 

g 

6 

© 

O 

Alcohol, 90% 
by Volume. 

© 

r- 

d 

< 

Brown Coal, 
or Lignite. 

oa 

M 

0 

o 

3 

o 

1 

6 

Wt. of 1 cu.ft., lbs. . 
Vol. of 1 lb., cu.ft. . . 

.032 

31.2 

.056 

17.9 

.068 

14.7 

.077 

13.0 

49.8 

.0201 j 

51.7 

.0194 

43.1 

.0232 

52.0 

.0193 

49.5 

.0202 

43.5 

.0230 

27.8 

.0360 

19.8 

.0507 

Heating value, 
B.T.U. per lb. . . . 
B.T.U. per cu.ft. . 

17450 

560 

17800 

995 

1940 

132 

1370 

105 

18900 

942000 

18000 

928000 

19700 

850000 

10000 

519000 

11700 

580000 

5400 

235000 

12600 

350000 

5000 

99000 

Vol. per 1000 B.T.U., 
ff, . 

1.79 

1.005 

7.60 

9.54 

.00106 

.00108 

.00118 

.00193 

.00173 

.00426 

.00286 

.0101 

Heat density (kero¬ 
sene =100) . 

.06 

.11 

.014 

0.11 

100 

| 98 

90 

55 

62 

25 

37.5 

11 

Cost of 1 lb., cents .. 

3.12 

2.68 



1.97 

.73 

2.62 

5.80 

.25 

.10 

.20 

.38 

Cost of 1000 B.T.U., 

.179 

.151 



.104 

.041 

24400 

.133 
i 7500 

.580 

.0214 

.0185 

.0158 

.0760 

cents . 



1 7 QO 

46800 

54000 

63200 

13200 

B.T.U. for 1 cent.. . 

5600 

6600 



9600 

I /oU 



















































514 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 

The above table is based on the following schedule of costs: Illuminating gas, 
$1.00 per 1000 cu.ft.; oil gas, $1.50 per 1000 cu.ft.; kerosene, 13 cents per gallon; 
alcohol, 40 cents per gallon; anthracite, $5.00 per ton; lignite, $2.00 per ton; coke, 
$4.00 per ton; wood, $3.00 per cord of 40 cu.ft. (8 ft. X4 ft.Xl5 in.). 


I. Fuel Gases 

For description of gas producer installations, seep. 268; for theoretical discussions 
of the gasification process, see Appendix, Section 5. 

In regard to gas producing apparatus, gas-engine builders confine their attention 
to the gasification of solid fuels; where illuminating gas is used in an engine, it is 
nearly always made in independent plants by the usual distillation process. The 
reason for this is that the various kinds of producer gas are very easily made and 
cleaned with the simplest Of apparatus. The manufacture of oil gas requires apparatus 
somewhat less complicated or costly than that used for illuminating gas, but the 

former gas, as far as its specific cost of heat is concerned, can only compete with 
the latter when the raw materials from which it is made are available at very low 
prices . 1 On this account oil gas, in spite of its great heating value and purity, 
generally finds application in engines only in places where it is used at the same 

time for some other and more important purpose. 

1. Illuminating Gas. Illuminating gas is produced by the dry distillation of 
bituminous coals in closed retorts, which are heated to bright redness. Under these 
conditions the volatile parts of the coal are driven off, while the by-products are coke, 
ammonia, and a small quantity of graphite. Each 100 lbs. of average coal will yield 
from 400 to 480 cu.ft. of cooled gas, from 50. to 70 lbs. of coke, 4.75 lbs. of tar and 
from 8 to 10 lbs. of ammonia liquor having a specific gravity varying from 1.5 to 
2.0° B. The heating of the retorts in an average plant requires from 15 to 20 lbs. 

of coke for every 100 lbs. of coal distilled. Allowing for the heating value of the 

coke so, used, we find that of the original amount of heat contained in the coal, 

20 % is found in the gas and from 50 to 60% is recovered in coke and tar. The 

manufacturing cost of the gas of course varies with the size, arrangement and operation 
of the plant and with the cost of coal. The sale price usually ranges from $1.00 
to $1.50 per 1000 cu.ft. 

The composition of the gas is subject to constant change, even in the same plant. 
Its main constituents are hydrogen (45-48% by vol.), methane (35-38%), carbon 
monoxide (5-8%), and several per cent, of heavy hydrocarbons, carbon dioxide, and 
nitrogen. A number of gas analyses are given in Table 119. The contents of so- 

called illuminants in the gas has considerable influence on the heating value of the 

gas, but illuminating power and heating value do not bear to each other a sufficiently 
definite ratio to enable the determination of the latter on the basis of the former, 
as used to be done. The heating value varies from 450 to 680 B.T.U. per cu.ft., 
and in some isolated instances may reach 780 B.T.U. Since the introduction of 
Welsbach mantles and similar devices and owing partly also to the increasing cost 
of coal, the average heating value of the illuminating gas has shown a considerable 
decrease in the last few years. At present the mean value is probably not far from 
560 B.T.U. per cu.ft. The best means of determining it is the calorimeter, but where 


1 This does not refer to the Lowe oil gas. 



FUELS 


515 


the composition of the gas is very definitely known, it may be found with almost 
equal accuracy by computation. 

The density of illuminating gas varies from .35 to .45, (air=1.0), the specific 
weight is on the average .032 lb. per cu.ft. Gas holders are usually kept under a 
pressure of from 4 to 8 ins. of water. From these the gas enters the mains through 
the station pressure regulator under a pressure varying from 2 to 3 ins. of water, 
while the pressure in the most distant branch lines is in most cases only .8 to 1.2 
in. For that reason it is usual to test out illuminating gas engines in the factory 
with a maximum gas pressure not to exceed .8 in. of water. The temperature of 
the gas in the mains is nearly always several degrees lower than the temperatUie 
of the outside air in summer, while in winter it may under certain circumstances 
be considerably higher than the air temperature. In mains put down rather deep 
the gas temperature is never found outside of the limits of 40 to 60° F., except on 
very cold or very hot days. 


Table 119 


ANALYSES OF ILLUMINATING GASES 



City. 

Methane. 

CHi. 

Hydrogen. 

H. 

Carbon 

Monoxide. 

CO. 

Carbon 

Dioxide. 

C0 2 . 

Oxygen. 

O. 

Heavy 

Hydro¬ 

carbons, 

^nlbn- 

Nitrogen, 

(Re¬ 

mainder). 

N. 

1 

Berlin (coal from Upper Silesia). . . 

32.7 

49.7 

9.5 

2.5 


4.6 

1.0 

2 

Konigsberg (Blochmann). 

36.5 

49.0 

5.6 

1.1 


6.8 

1 .0 

3 

Magdeburg (Gasjourn. 1900, 87) .. . 

30.1 

54.9 

7.7 

1.4 

0.2 

3.3 

2.4 

4 

Dresden (Schottler). 

33.4 

48.7 

8.0 

1 .5 

1 .4 

3.0 

4.0 

5 

Hannover (Schottler). 

37.5 

46.3 

11.2 

0.8 


3.2 

1.0 

6 

Frankfurt a. M. (Leybold). 

32.6 

49.8 

8.8 

2.3 


4.0 

2.5 

7 

Bonn (Clerk). 

43.1 

39.8 

4.7 

3.0 


4.7 

4.7 

8 

Heidelberg (Bunsen). 

34.0 

46.2 

8.9 

3.0 

0.6 

5.1 

2.2 

9 

Aachen (v. Ihering). 

34.2 

54.0 

5.2 

1.1 


3.3 

2.2 

10 

Paris (Peclet). 

33.1 

50.1 

6.3 

1.5 

0.5 

5.8 

2.7 

11 

London (Gaslight & Coke Co.) . . . . 

37.6 

48 1 

3.7 


0.3 

4 .4 

6.0 

12 

Manchester (Clerk). 

34.9 

45.6 

6.6 

3.7 


6.5 

2.7 


Average. 

34.99 

48.51 

7.183 

1 .825 

0.250 

4.560 

2.700 


The heating value of No. 1 was approximately 560 B.T.U. per cu.ft.; of No. 3, 
600 B.T.U. per cu.ft., and of No. 9, 628 B.T.U. per cu.ft. 


Table 120 


CONSTANTS FOR ILLUMINATING GAS 


Composition of Illuminating Gas. 

1 Cubic Foot of Gas Contains 

Hu in B.T.U. 

Air Theoretically Required. 

Cubic Feet. 

Pounds. 

Cubic Feet. 

Pounds. 

Hydrogen . H. . 

Methane . CH 4 . 

Carbon monoxide.C'O.. 

Heavy hydrocarbons.0 2 H 4 

Carbon dioxide.C0 2 • 

.4850 
.3500 
.0700 
.0450 
.0200 
.C 025 
.0275 

.00271 

.01562 

.00546 

.00351 

.00245 

.00022 

.00215 

144.0 

333.2 

23.9 

70.4 

1.145 

3.285 

.165 

0.639 

- .009 

.0924 

.2651 

.0133. 

.0516 

- .0007 

Referred to 1 cu.ft. of gas. 

1.0000 

.03212 

571.5 

5.225 

.4217 

























































516 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


Table 120 contains the main constants for illuminating gas of most interest to 
the designer, as computed from the average of the twelve analyses above given. 
Regarding the method of obtaining the figures in the various columns, Section 4, 
Appendix, and especially Tables 147 and 148 will give detailed information. 

Average Constant for Illuminating Gas: Specific weight, .032 lb. per cu.ft.; 
density, .4 (air =1.0); lower heating value, 560 B.T.U. per cu.ft., or 17 500 B.T.U. 
per lb., minimum air required, 5.25 cu.ft. per cu.ft., or .4217 lb. per cu.ft., or 13.2 
lbs. per lb. 

2. Oil Gas. Oil gas is made from cheap crude or industrial oils by vaporizing 
them in retorts at red heat and further heating the resulting oil vapor. The kind of 
oil used depends largely upon the locality where the gas is made. German oil gas 
installations for instance use almost entirely a paraffine oil which is a by-product of 
the manufacture of paraffine from bituminous coal tar; in other places suitable crude 
oils or petroleum residuum, when the latter can be had in sufficient quantity, are 
employed. 100 lbs. of gas oil will develop from 750 to 950 cu.ft. of cooled gas having 
a specific weight of from .043 to .058 lbs. per cu.ft. and a lower heating value varying 
from 900 to 1100 B.T.U. per cu.ft. The amount of coke used in heating the retorts 
for every 100 lbs. of oil gasified is from 70 to 80 lbs. The process therefore recovers 
from 30 to 40% of the heat contained in oil and coke. 

The composition of oil gas depends mainly upon the kind of oil from which it 
is made. The principal constituents are always hydrocarbons, from which the gas 
also obtains its high heating value. 1 According to the firm of Julius Pintsch, Berlin, 
an oil gas made from residuum of petroleum gave the following volumetric analysis: 
17% ethylene, 58% methane, and 24.3% hydrogen. Another oil gas, made from 
paraffine oil, contained 28.9 volume per cent of ethylene, 54.9% methane, 5.6% hydrogen, 
8.9% carbon monoxide, and .9% carbon dioxide. Table 121 following is based upon 
the latter analysis, the remainder of which is not known. 

Table 121 


CONSTANTS FOR OIL GAS 


Composition of Oil Gas. 

1 Cubic Foot of Gas Contains 

H u B.T.U. 

Air Theoretically Required 

Cubic Feet. 

Pounds. 

In Cubic Feet. 

In Pounds. 

Hydrogen. 

.H. . . 

.056 

.0031 

16.63 

.132 

.0106 

Methane. 

•ch 4 . 

.549 

.0245 

522.64 

5.152 

.4158 

Ethylene. 

. ,c 2 h 4 

.289 

.0226 

451.99 

4.110 

.3317 

Carton monoxide. 

. .CO. . 

.089 

.0069 

30.43 

.021 

.0017 

Carton dioxide. 

..CO,. 

.009 

.0011 




Referred to 1 cu.ft. of gas. 

.992 

.0583 

1021.7 

9.415 

.7598 


Average Constants for Oil Gas: Specific weight, .058 lbs. per cu.ft.; density, .6 
(air = 1.0); lower heating value 1000 B.T.U. per cu.ft. or 17 300 B.T.U. per lb.; 
minimum air required, 9.5 cu.ft. per cu.ft., or .76 lb. per cu.ft., or 13.2 lbs. per lb. 


1 This type of oil gas should be distinguished from the water-oil gas made by the Lowe or similar 
processes. 

























FUELS 


517 


3. Power or Producer Gas. Producer gas burns with a blue (non-illuminating) 
flame and is a modified water gas used mainly for power or heating purposes. It is 
manufactured in generators usually forming a part of the engine installation. The 
process consists in leading a mixture of air and steam through an incandescent layer 
of coke or coal, the result being that the carbon dioxide produced by the com¬ 
bustion is reduced to carbon monoxide, that the steam is decomposed, forming hydrogen 
and carbon dioxide, and finally the latter gas is also largely reduced to carbon 
monoxide. The first, and for some years only successful, power gas generator was 
developed by the English engineer Dowson, 1 from whom this gas derived the name 
by which it is largely known even to-day. 

The most suitable fuels for generators are gas coke and coals which do not 
contain any easily condensible gases and as little as possible of any non-combustible 
ingredients. These fuels give little trouble by interfering with continuous operation 
and furnish a gas that may be easily cleaned. Coke and anthracite coal are mostly 
used, and only very recently have attempts to build generators and cleaning apparatus 
for bituminous coal met with any considerable measure of success. 

One pound of coke or anthracite will yield from 65 to 77 cu.ft. of gas having 
an average heating value of from 125 to 135 B.T.U. per cu.ft. The amount of water 
used per lb. of fuel is generally in the ratio of 1 to 1. A water supply greater than 
this is not economical, considered from the standpoint of thermal efficiency (see p. 
521). The temperature at which the gas leaves the generator first decreases and 
then increases between, any two charging periods. The range of variation is between 
900 and 1350° F. The density of the gas varies from .85 to .95 (air =1.0), so that 
its specific weight is from .068 to .076 lb. per cu.ft. The combustion of 1 cu.ft. 
of this gas requires theoretically from .95 to 1.1 cu.ft. of air. Being a mixture of 
air-gas and water-gas, producer gas consists mainly of carbon monoxide, hydrogen, carbon 
dioxide and nitrogen. The composition, however, depends largely upon the existing 
condition in the generator, and is therefore subject to constant and considerable 
change. Table 122, taken from the report of a test made by E. Meyer , 2 clearly 
shows this. 

Table 122 

PRODUCER GAS ANALYSES 



Time of Taking Gas Sample. 

Mean of 
the 
Eight 
Samples 

9:40 

10: 55 

11:55 ; 12:56 

2:04 

3:15 

4:20 

8:10 

fC0 2 

I CO 

Composition in volume, %. J CH, 

1 H 

O 

L (diff.) N 

6.5 

26.6 

1.3 

6.8 

.1 

58.7 

4.8 
28.2 

2.7 

5.9 

0 

58.4 

4.2 

27.8 

2.6 

6.4 

0 

59.0 

4.7 

26.7 

1.8 
8.8 

.1 

57.8 

5.0 

26.6 

2.2 

9.1 

0 

57.1 

4.0 

29.0 

1.7 

8.0 

0 

57.3 

4.8 
28.4 

2.0 

5.8 
.2 

58.8 

4.6 

27.8 

1.6 
5.1 

.1 

60.9 

4.8 

27.6 
2.0 
7.0 

0 

58.6 


Soon after each charging period the percentage of combustible gases always 
considerably increases. In some cases the gases of distillation from the fresh fuel may 


1 See German patent, No. 27165, 1887. 

2 Z. d. V. D. I., 1896, p. 1239 and 1304. 























518 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


even double the normal heating value of the gas, especially if a considerable decrease in 
the volume yield of the producer is combined with the charging period. Of the total 
fuel used in pressure gas installations, from -fa to £ is used in separate boilers for 
the making of the necessary steam. In suction-gas plants the heat required for this 
purpose is furnished by the producer itself. Through incomplete combution, direct 
loss of gas and radiation, from 20 to 25% of the heat developed in the producer is 
lost; so that from 75 to 80, in specially favorable cases up to 83%, of the heating 
value of the fuel is recovered in the gas. Owing to stand-by losses this efficiency 
is decreased to from 60-70%. 

The mathematical determination of the amount of gas produced per pound of 
coal and the resulting efficiency of the gasification process is treated in Section 5 of 
the Appendix. Prof. E. Meyer, in the report above mentioned, used the following 
method of computation which, on account of the resulting simple equations for V 
and rj t , serves to give a rapid numerical survey of the entire process. 


Assume that the chemical analysis of the gas shows the following: a vol. per cent C0 2 , 
b vol. per cent CO, and c vol. per cent CH 4 . It is also assumed that the gas contains no other 
hydrocarbons. 

Now two volumes of C0 2 were produced from two volumes of oxygen and one volume of 
carbon, the latter as a gas having a specific weight twelve times that of hydrogen, that is, 
12X.00559 = .067081 lb. per cu.ft. at 32° F. and 760 mm. barometer. Hence in 1 cu.ft. of C0 2 
there must be I cu.ft. of carbon (considered as gas) or ^X.06708 = .03354 lb. of carbon. 

The same amount of carbon is, however, also found in 1 cu.ft. of CO and in 1 cu.ft. of 
CH 4 , as similar computation will show. 100 cu.ft. of gas of the composition above given there¬ 
fore contain .03354 (a+b+c) pounds of carbon, or 1 pound of carbon produces 


V' = 


100 


2981 


.03354(o + 6-l-c) a+ b+c 


cu.ft. of gas , 


( 1 ) 


at 32° F. and 760 mm. barometer. 

The fuel, however, never consists of pure carbon. One pound of the former contains <f> 
pounds of the latter, in which <£<1>0 and is determined from a chemical analysis of the fuel. 
One pound of fuel will then produce 


r 2981 
a+b+c 


cu.ft. of producer gas, 


( 2 ) 


irrespective of what ingredients other than the carbon combinations above specified the gas may 
contain. If, however, the gas, besides, C0 2 , CO, and CH 4 , also contains C 2 H 4 , so that the 
analysis shows, say d volume per cent of C 2 H 4 , the production of gas per pound of fuel then 
becomes 


v 2981 
a + 5+c-f-2d 


cu.ft., 


(2a) 


since 1 cu.ft. of C 2 H 4 contains 2 X .03354 = .06708 lb. of C. 

Let H be the lower heating value of 1 cu.ft. of producer gas, and K the lower heating 
value of 1 lb. of the fuel. Of the latter, the part B g lbs. is used in the generator and the part 









FUELS 


519 


B d in the boiler, the ratio - being assumed as known. Further, if is the thermal efficiency 
of the entire producer plant, then 


VB U H VII 

K(B,j + Bd) ^ 1 + Bj\j g . 


The ratio — may for approximate computations be put equal to .1 for installations having 

a separate boiler, where no separate boiler is employed, |^=0. In the former case we may 

then write 


2981#/ 
l.l(a + 6 + c)A 


(4) 


If h represents the percentage of hydrogen in the gas by volume, the heating value of 
1 cu.ft. of the gas produced may be expressed by 


A =(3.426 +9.52c+ 2.97/0 B.T.U. 


(5) 


The computation, however, will give more accurate results if H is determined by means of 
the Junker’s calorimeter instead of by computation based upon the chemical analysis. 

In the case under discussion (according to Table 122), we have that a + b + c = 34.4, hence 


29S1 

U' = —= S(5.7 cu.ft., 


that is, 1 lb. of carbon gasified produces 86.7 cu.ft. of producer gas. 

Further, <£ = .8772, which makes U = 76.1 cu.ft., i.e., 1 lb. of fuel produces 76.1 cu.ft. of gas, 
measured at 32° F. and 760 mm. It was also found that 


!?</ = .109 B a , K= 13 000 B.T.U. per lb. of coal; 

(H)' = 134.0 B.T.U. per cu.ft., as computed from the analysis; or, 
H = 135.0 B.T.U. per cu.ft., by calorimeter determinations. 


With these figures, 


and 




76. 1X134 

1.109X13000 


m= 


76.1 X135 _ 


1.109X13000 


The formula just established for Vt is valid to the fullest extent only when all of the fuel 
charged is really gasified and when no fuel particles, which are not afterward recovered are 













J 


520 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 

removed with the ash. If we let B a represent that part of the fuel lost in the ash, the 
expression for the efficiency becomes 


V(B„-Ba)H 
K(B g + B d ) ’ 

The constants for producer gas in the following table have been computed on 
the basis of the average analysis from Table 122. 1 


Table 123 

CONSTANTS FOR COKE PRODUCER GAS 


Composition of Producer Gas. 

1 Cubic Foot of Gas Contains 

H u in B.T.U. 

Air Theoretically Required 

Cubic Feet. 

Pounds. 

In Cubic Feet. 

In Pounds. 

Hydrogen. 

.. .H. . 

.070 

.0004 

20.8 

.165 

.0133 

Methane. 

. ,ch 4 

.020 

.0004 

19.0 

.188 

.0152 

Carbon monoxide. 

..CO . 

.276 

.0215 

94.2 

.649 

.0525 

Carbon dioxide. 

• CO,. 

.048 

.0058 




Nitrogen. 

. .N . . 

.586 

.0458 




Referred to 1 cu.ft. of gas. 

1.000 

.0744 

134.0 

1.002 

.0810 


Average Constants for Coke Producer Gas: Specific weight, .075 lb. per cu.ft.; 
density, .93 (air =1.0); lower heating value, 135 B.T.U. per cu.ft., or 1800 B.T.U. 
per lb.; minimum air required, 1 cu.ft. per cu.ft., or .081 lb. per cu.ft., or 1.10 lbs. 
per lb. The above gas was made from coke having a heating value of 13 210 B.T.U. 
and therefore' contains but little hydrogen, which manifests itself in the low heating 
value and high specific weight of the gas. Producer gas from anthracite always 
contains more hydrogen, consequently shows a higher heating value and lower weight. 
The difference is clear from Table 124, which gives the constants for a producer gas 
made from a Belgian hard coal having a heating value of 14 200 B.T.U. 2 

Table 124 


CONSTANTS FOR ANTHRACITE PRODUCER GAS 


Composition of Producer Gas. 

1 Cubic Foot of Gas Contains 

H u in B.T.U. 

Air Theoretically Required 

Cubic Feet. 

Pounds. 

In Cubic Feet. 

In Pounds. 

Hydrogen. 

. H. . 

.242 

.0014 

71.8 

.572 

.0462 

Methane. 

• CH 4 

.020 

.0009 

19.0 

.188 

.0152 

Carbon monoxide. 

. .CO. 

.166 

.0129 

56.7 

.390 

.0314 

Carbon dioxide. 

. co ? . 

.113 

.0139 




Nitrogen. 

. . N . . 

.459 

.0359 




Referred to 1 cu.ft. of gas. 

1 .000 

.0650 

147.5 

1.150 

.0928 


l In conjunction with Tables 147 and 148, Appendix. 

2 Analysis of coal is given in Z. d. V. D. I., 1896, p. 350, of the gas in Z. d. V. D. I., 1895, p. 1540. 














































FUELS 


521 


Average Constants for Anthracite Producer Gas: Specific weight, .065 lb. per 
cu.ft.; density, .8 (air=1.0); lower heating value 147 B.T.U. per cu.ft. = 2260 B.T.U. 
per lb.; minimum air required, 1.15 cu.ft. per cu.ft., or .093 lb. per cu.ft., or 1.45 
lbs. per lb. 

An ample supply of steam to the generator is of advantage from a practical 
standpoint, since it tends to decrease clinkering and to prevent the rapid burning, 
away of lining and grates. Too high a percentage of hydrogen in the gas, however, 
leads to heavy explosions in the cylinder of the engine. Only a few engines can stand 
from 7 to 10% of hydrogen in the mixture, i.e., from 15 to 20% in the producer gas;, 
in most of them, under continued heavy load, a troublesome knocking appears as soon 
as the gas contains more than 10% of hydrogen. The composition of the producer 
gas should therefore not be made entirely dependent upon the efficiency of the gasifi¬ 
cation process (see p. 536 and 597). 

4. Blast Furnace Gas. Blast furnace gas, although a gas low in heating value, 
is yet readily combustible, and, forming a by-product of blast furnace operation, is- 
available nearly withput cost in such quantities that one-fourth of it would be suffi¬ 
cient to cover the power demand of the blast furnace, leaving the other three-fourths 
available for other purposes. The manufacturer of 1 ton of pig iron requires the use 
of roughly 1 ton of coke together with about 4 tons of air (blast). The production 
of 1 ton of pig iron is therefore accompanied by the evolution of 5 tons, that is 
140 000 cu.ft., of blast furnace gas, and a medium sized stack of say 150 tons capacity 
therefore furnishes about 21 000 000 cu.ft. of gas in 24 hours. Assuming that one-half 
of this volume is used in the hot blast stoves of the plant, this leaves 10 500 000 
cu.ft. available, which quantity used in gas engines is sufficient to generate at least 4000 
H.P. If used under steam boilers the same volume of gas would not produce 1500 engine 
H P Realizing that some blast furnace plants produce from 1100 to 1200 tons of 
pig iron every 24 hours, that is, eight times the capacity assumed, the great import¬ 
ance as far as the steel works are concerned, of the direct utilization of blast furnace 
gas as fuel gas in engines becomes at once apparent. To Fr. W. Lurman belongs the 
credit of first having pointed out, in 1886, the availability of blast furnace gas for 
use in gas engines, and to have induced gas engine builders into experimenting 


with this gas. . , 

Since the blast furnace is fundamentally nothing but an immense gas producer, 

it will suffice to refer to page 592 of the Appendix regarding the process of gasifi¬ 
cation. In this case, however, the charge not only consists of coke but m addition 
also of the burden, that is of certain quantities of iron ore and of flux, usually 
limestone. This modification makes both the carbon dioxide and the water vapor 
content of the resulting gas greater than the percentage of the co ™ s P on * ng , 
uents found in ordinary producer gas. Neglecting the large percentage of dust and 
water vapor carried, blast furnace gas contains about the same component gases as 
producer gas, that is mainly carbon monoxide and nitrogen and, as secondary 
admixtures carbon dioxide and hydrogen. The composition of the gas, owing to its 
manufacture as a by-product, varies considerably in different*3 ^^^ 
different intervals. The combustible constituents fluctuate between 25 and M%, and 

^tb^bfi^S LUrman, Ogives some detailed information regarding 
this point. 


1 Stahl und Eisen, Heft 21, 1901. 



522 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


Table 125 

I 

ANALYSES OF BLAST-FURNACE GASES 


Composition of Blast-furnace Gas. 

Eisenh ut tenkunde 
von Ledebur, 1893 

p. 100. 

Works in West¬ 
phalia. 

Works in 

Average. 

8 

Dry. 

1 

Containing 
10% H 2 O. 
2 

Dry. 

3 

Containing 
10% H 2 O. 
4 

Upper 

Silesia. 

5 

Upper 

Silesia. 

6 

The 

Minette 

District. 

Carbon monoxide.. . . 

.. .CO 

24.0 

21.6 

29.0 

26.1 

29.7 

23.0 

27.5 

25.84 

Hydrogen. 

.. .H 

2.0 

1.8 

4.0 

3.6 

6.3 


3.0 

2.96 

Methane. 

... CI1 4 

2.0 

1.8 






.54 

Carbon monoxide.. . . 

...C0 2 

12.0 

10.8 

10.0 

9.0 

7.8 

6.0 

10.0 

9.37 

Nitrogen. 

.. .N 

60.0 

54.0 

57.0 

51.3 

56.2 

59.0 

54.5 

56.00 

Water vapor. 

. . H 2 0 


10.0 


10.0 


12.0 

5.0 

5.29 

X in lbs./cu.ft. 


.0814 

.0782 

.0795 

.0767 

.0799 

.0775 

.0790 

.0789 

H u in B.T.U./cu.ft. . . 


105.7 

95.3 

109.6 

98.7 

100.0 

77.5 

101.7 

98.5 


The greatest, at first apparently insurmountable, obstacle that stood in the way 
of the successful utilization of blast-furnace gas in the gas engine was the large 
amount of dust carried by it. In especially unfavorable cases 1 cu.ft. of gas may 
carry .6 gram of dust, besides certaining metallic vapors which assume the form of 
dust only after combustion in the cylinder. It is comparatively easy to take out of 
the gas the coarse coke, iron oxide or lime stone dust, for which purpose even the simple 
cleaning apparatus installed in many works before the introduction of the gas engine 
is quite sufficient. These cleaners or washers are, however, not able to take care of 
the fine dust, which is of most danger to the engine, and it therefore became 
necessary first to invent and construct suitable apparatus. Although this important 
problem is as yet not completely solved, it may be said that the present-day 
scrubbers are capable of reducing the dust content of very dirty gas to such an extent 
that neither the reliability of operation or the durability of the blast furnace gas 
engine need any longer be seriously questioned. 

To thoroughly clean the gas of the very fine dust which remains suspended in 
the gas current for hundreds of feet, is only possible by means of some wet method, 
which accordingly is the scheme most commonly adopted. After the gas has been given 
a preliminary dry cleaning by passing it through settling-tanks, dust-catchers, etc., it is 
next led through scrubbers, wet blowers or ventilators and other types of washers 
and, emerging from these, it is led directly to the engine or to gas holders. The 
large quantity of water vapor and the metallic vapor that the gas originally carries 
are also largely thrown down in these wet scrubbers. The water used in the latter 
varies from 2 to 10 cu.ft. per 1000 cu.ft. of gas cleaned. The quantity required is 
therefore considerable under certain circumstances, and for that reason it is in many cases 
recovered for future use by means of settling tanks or cooling ponds. Coke scrubbers 
for blast furnace gas have not been very successful. The passages between the 
layers of coke and between the individual pieces are soon partly or entirely stopped 
up by the coarse dust, and the flow of gas gradually ceases. A number of blast 
furnace gas cleaning installations in actual operation have been described in detail 


























FUELS 


523 


by Liirman and others, and the figures in the following table have been abstracted 
from these descriptions. 


Table 126 

FIGURES RELATING TO THE OPERATION OF BLAST-FURNACE GAS-CLEANING APPARATUS 


1. Dust content of gas: 

(а) Ahead of the dry cleaner, in 1 cu.ft. of 

gas, grams. 

After passing dry cleaner, in 1 cu.ft. of 
gas, grams. 

( б ) Ahead of the wet scrubber, in 1 cu.ft. of 

gas, grams. 

After passing wet scrubbers, in 1 cu.ft. 
of gas, grams. 

(c) Ahead of last cleaner, in 1 cu.ft. of 

gas, grams. 

After passing last cleaner, in 1 cu.ft. of 
gas, grams. 

( d) At the engine. 

12. Water carried by gas: 

(a) Ahead of the cleaning app., in 1 cu.ft. of 

gas, grams. 

(b) After passing cleaning app., in 1 cu.ft. of 

gas, grams. 

3. Temperature of gas: 

(a) Ahead of cleaning apparatus, 0 F. 


(b) After passing cleaning apparatus, ° F. . .. 

(c) At the engine, ° F. 

4. Pressure, inches of water: 

(a) Ahead of cleaning apparatus, in. .. 

(b) After passing cleaning apparatus, in. 


■5. Quantity of gas washed per hour, cu.ft.• 


h. Quantity of washing or cooling water, per hour, 
cu.ft. 

7. Quantity of dust removed per hour, lbs. 

8 . Consumption of scrubber water: 

(a) Per eff. H.P. hour, cu.ft. 

(b) Per cubic foot of gas, cu.ft. 

9. Cooling or settling ponds: 

(a) Superficial area, sq.ft. 

(b) Depth of water, in. 

(c) Cost of construction, dollars. 


Gute 

Hoffnungs 

Hiitte. 

Georgs- 

Marien 

Hiitte. 

Friedens- 

Hiitte 

O.-S. 

Dudelin- 

gen. 

Donners- 

markhutte 

.167 

1 

.212 

.142 


.... 

\ .150 





J 



.137 

.... 

.085 

.083 

.017-.045 

.011 




f 

.... \ 

.007 

1 

.013 


behind 

blower 

| .071 

1 

.007 


.0006 

.005 

.003 

7 vol. % 

4.02 

.382 


.850 

/ 1.5 

l vol.% 

.772 

.156 


.056 

r 1 

338 

323 

262 

194-212 

102 

105-113 

75 

46-55 

48-53 

63 J 

air-temp. 

.... 


27-53 

1 

5.92 


3.56-5.92 

. 2 -.4 

.21 

3.56 


.78-2.36 

1.97 

1.97 

925000 

847000 



77500 

88000 


for stoves 


353000 

to 

to 


78000 for 



38800 

53000 


engines 





762 

1765 

3530 


-•( 

.012 cu.ft, 
per cu.ft. 

132.5 


18.8 

.... 


.236 



.... 


.067 

.353 


.... 

.... 

465 

1290000 

5380 

.... 

/ 

43.4 

59.0 


.... 


23400 

7500 

34500 


2700 


Differ- 

dingen. 


302 

151 

077 

Oil 


Ahead of 
blower 
115 
After 
passing 
blower 
105 
73-82 

1.18 

3.16-3.96 


318000 

495 

78.0 

.166 

.055 

12700 

cu.ft. 


It will be seen that the quantity of dust the gas may carry is extremely variable. 
It sometimes happens that the gas at the outset only contains from .007 to .014 












































524 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


gram of dust per cu.ft., in which case wet cleaning may be dispensed with altogether. 
According to the Societe Cockerill, Seraing, however, wet scrubbing is necessary in all 
cases where haematite, or mixtures of the same, are reduced in the blast 
furnace. 

The following table of constants for blast furnace gas is based upon the last 
column of Table 125, the analysis there given having been approximated to some 
extent without great error. 


Table 127 

CONSTANTS FOR BLAST FURNACE GAS 


Composition of Blast Furnace Gas. 

1 Cubic Foot of Gas Contains 

H u in B.T.U. 

Air Theoretically Bequired 

Cubic Feet. 

Pounds. 

In Cubic Feet. 

In Pounds. 

Hydrogen . 

H 

.030 

.0002 

8.9 

.071 

.0057 

Methane. 

• ch 4 

.005 

.0002 

4.7 

.005 

.0004 

Carbon monoxide. 

.CO 

.260 

.0203 

88.8 

.611 

.0493 

Water vapor. 

h 2 o 

050 

.0025 




Carbon dioxide. 

.co 2 

.095 

.0116 




Nitrogen. 

.N 

.560 

.0438 




Referred to 1 cu.ft. of gas. 

1.000 

.0786 

102.4 

.687 

.0554 


Average Constants for Blast Furnace Gas: Specific weight, .0786 lb. per cu.ft.; 
density, .98 (air = 1.0); lower heating value, 101 B.T.U. per cuit., or 1290 B.T.U. per 
lb.; minimum air required, .7 cu.ft. per cu.ft., or .055 lb. per cu.ft., or .72 lb. per 
lb. 

5. Coke Oven Gas. Coke oven gas is another of the industrial gases, being a 
by-product of the manufacture of coke used for metallurgical purposes (dry distillation 
of bituminous coal). There is fundamentally no difference between this process and 
the manufacture of illuminating gas. Coke oven gas therefore possesses the same 
heating value and. in general much the same characteristics as illuminating gas. It 
is for these reasons a good fuel for the operation of gas engines, and, after a period 
of senseless waste which lasted, until a very short time ago, this gas is now used in 
many places in gas engines with great economy. As it appears at present, the coke 
oven gas engine has a successful future before it. 

According to the statements of Schmidt, in “ Stahl und Eisen,” 1901, p. 259, 
1000 lbs. of bituminous coal will in coke ovens yield 3800 cu.ft. of coke oven gas, of 
which approximately one-third is first recovered as illuminating gas, while the remaining 
two-thirds are used for power and heating purposes. Table 128 is based upon the figures 
resulting from averaging a large number of analyses of coke oven gas, which have 
appeared in “Stahl und Eisen” and in the Gas-Journal. The heating value, according to- 
the figures available, varies between 645 and 338 B.T.U. per cu.ft. 

Average Constants for Coke Oven Gas: Specific weight, .027 lb. per cu.ft., 
density, .36 (air =1.0); lower heating value 545 B.T.U. per cu.ft., or 18 700 B.T.U. 
per lb.; minimum air required, 5 cu.ft. per cu.ft., or .4 lb. per cu.ft., or 14.8 lbs- 
per lb. 























FUELS 


525 


Table 128 

CONSTANTS FOR COKE OVEN GAS 


Composition of Coke Oven Gas. 

1 Cubic Foot of Gas Contains 

H u in B.T.U. 

Air Theoretically Required 

Cubic Feet. 

Pounds. 

In Cubic Feet. 

In Pounds. 

Hydrogen. 

.H 

.550 

.0031 

163.0 

1.299 

.1045 

Methane. 

.ch 4 

.320 

.0143 

304.5 

3.008 

.2420 

Ethylene. 

.c 2 h 4 

.015 

.0012 

23.4 

.213 

.0172 

Benzol . 

• C e H e 

.008 

.0018 

31.0 

.284 

.0229 

Carbon monoxide. 

.CO 

.070 

.0055 

24.0 

.165 

.0133 

Water vapor. 

. HD 

.010 

.0005 




Carbon dioxide. 

.co 2 

.012 

.0015 




Nitrogen. 

,N 

.015 

.0012 




Referred to 1 cu.ft. of gas. 

1.000 

.0291 

545.9 

4.969 

.3999 


6. Brown Coal or Lignite Gas. This gas is apparently cf no importance as 
far as the United States are concerned. It is one of the by-products of the manu¬ 
facture of tar from brown coal, the others being coke and ammonia. Its composition 
is a good deal like that of coke oven gas, but it is considerably lower in heating 
value. 1 cu.ft. of brown coal will yield about 120 cu.ft. of gas. The latter is readily 
combustible in gas engines, and has been used foi this purpose for years wherever 
available. It is not likely, however, that the gas will ever assume any great importance, 
since the manufacture of brown coal tar is a highly localized industry and not 
carried on to any extent. 

The figures in the table following are based upon data given by Dammer in his 
“ Chemical Technology.” 


Table 129 


CONSTANTS FOR BROWN COAL OR LIGNITE GAS 


Composition of Gas. 

1 Cubic Foot of Gas Contains 

H u in B.T.U. 

Air Theoretically Required 

Cubic Feet. 

Pounds. 

In Cubic Feet. 

— 

In Pounds. 

Hydrogen.H 

Methane.CH 4 

Heavy hydrocarbons.C n H 2n 

Carbon monoxide.CO 

Hydrogen sulphide.H 2 S 

Carbon dioxide. t - • -C0 2 

Oxygen.O 

Nitrogen.N 

.243 

.165 

.014 

.081 

.011 

.170 

.031 

.285 

.0014 

.0074 

.0011 

.0064 

.0010 

.0209 

.0028 

.0223 

72.0 

156.8 

27.7 

27.7 

.574 

1.550 
.199 
.190 

- .117 

.0462 

.1250 

.0160 

.0153 

- .0093 

Referred to 1 cu.ft. of gas. 

1.000 

.0633 

284.2 

2.396 

.1932 


Average Constants for Brown Coal or Lignite Gas: Specific weight, .065 lb 
>r cu.ft.; density, .80 (air =1.0), lower heating value 285 B.I.U per cu.it., or 4 0 
.T.U. per lb.; minimum air required, 2.4 cu.ft. per cu.ft., or .10 lb. per cu. ., or 

lbs. per lb. 




















































526 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


7. Acetylene. Since the commercial introduction of calcium carbide, acetylene,, 
a colorless hydrocarbon gas (C 2 H 2 ), has occasionally been used as a gas engine fuel. It 
is made from calcium carbide to which water is added at a uniform rate in simple 
gas producers. During the process the carbide breaks up into acetylene with the 
formation of calcium hydroxide as a by-product. Under favorable conditions, 1 lb. of 
carbide will yield 4.8 cu.ft. of acetylene. The heating value of the gas is 1480 B.T.U. 
per cu.ft. and the specific weight .074 lb. per cu.ft. Its density, referred to air, is 
therefore .91. For complete combustion this gas requires theoretically 11.85 cu.ft. 
per cu.ft., or .96 lb. per cu.ft., or 13.2 lbs. per lb. In practice the excess of air 
used should be very large to prevent the extremely severe explosions occurring with 
mixtures approximating the theoretical. Explosibility does not cease even with a 

Rir 

mixture in the volume ratio of — = 40. 

gas 

A general introduction of acetylene gas as a gas engine fuel is probably out of 
question on account of the high cost of carbide, the fire risks involved, and the 
unpleasant odor the gas possesses. 

8. Natural Gas. This gas is obtained from wells in various parts of the United 
States. Its composition differs somewhat according to locality, and sometimes also 
changes from time to time in the same well. In nearly all cases, however, the main 
constituent is marsh gas, CH 4 . The following table is based upon data given by Kent 
for a natural gas from Anderson, Indiana. 


Table 130 


CONSTANTS FOR NATURAL GAS 


Composition of Gas. 

1 Cubic Foot of Gas Contains 

H u in B.T.U. 

Air Theoretically Required 

Cubic Feet. 

Pounds. 

In Cubic Feet. 

In Pounds. 

Hydrogen. 

. . .H 

.0186 

.0001 

5.5 

.045 

.0036 

Marsh gas. 

• • ch 4 

.9307 

.0415 

885.9 

8.896 

.7117 

Olefiant gas. 

. . C 2 H 4 

.0047 

.0004 

7.4 

.068 

.0054 

Carbon monoxide. 

...CO 

.0073 

.0005 

2.5 

.016 

.0013 

Carbon dioxide. 

...co 2 

.0026 

.0003 




Oxygen. 

.. .0 

.0042 

.0004 


- .015 

- .0012 

Nitrogen. 

.. .N 

.0304 

.0024 




Hydrogen sulphide. 

.. ,h 2 s 

.0015 

.0001 




Referred to 1 cu.ft. of gas. 

1.000 

.0457 

901.3 

9.000 

.7208 


Average Constants for Natural Gas: Specific weight, .046 lb. per cu.ft.; density, 
.575 (air =1.0); lower heating value, 900 B.T.U. per cu.ft., or 19 600 B.T.U. per lb.; 
minimum air required, 9 cu.ft. per cu.ft. or .72 lb. per cu.ft., or 15.5 lbs. per lb. 

II. Liquid Fuels 

1. Crude Oil and its Distillates. The fuel oils are mainly used as fuels for 
the smaller sizes of internal combustion engines. Their use for larger sizes is, among 
other things, generally prohibited by their cost. Crude oil is a native, generally liquid, 
combination of various hydrocarbons, and as such usually consists of from 80 to 86%. 




























FUELS 


527 


of carbon and from 15 to 10% of hydrogen, which are the main constituents. There 
is usually also a small percentage of impurities. The relative proportions of the 
various constituents vary somewhat according to the locality where the oil is found, 
and also to some extent upon the oil’s geologic age, as is indicated in Table 131. * 1 


Table 131 

ANALYSES OF CRUDE OIL 



Specific 
Gravity at 

Composition. 

Heating 

Value 

B.T.U. 


32° 

r. 

Carbon, 

C. 

Hydrogen, 

H. 

Oxygen and 
Impurities. 

American petroleum. 

.820 

83.4 

14.7 

1.9 

17588 

Heavy crude, West Virginia. 

.873 

83.5 

13.3 

3.2 

18324 

Light crude, West Virginia. 

.841 

84.3 

14.1 

1.6 

18401 

“ Pennsylvania. 

.816 

82.0 

14.8 

3.2 

17933 

Heavy crude Pennsvlvania. 

.886 

84.9 

13.7 

1.04 

19210 

Crude oil Parma. 

.786 

84.0 

13.4 

1.8 

18218 

‘ 1 Pechelbronn. 

.912 

86.9 

11.8 

1.3 

17474 

(( < ( 

.892 

85.7 

12.0 

2'.3 

18036 

‘ ‘ Schwabweiler .. . 

.861 

86.2 

13.3 

.0 

18844 

i e < < 

.829 

79.5 

13.6 

6.9 


‘ ( Hannover. 

.892 

80.4 

12.7 

6.9 


ti ‘ < 

.955 

86.2 

11.4 

2.4 


(( Eastern Galicia. 

.870 

82.2 

12.7 

5.7 

18153 

Western Galicia . 

.885 

85.3 

12.6 

2.1 

18415 

Balapha/nv . 

.882 

87.4 

12.5 

.1 

21060 

Light crude Baku. 

.884 

86.3 

13.6 

.1 

20628 

TTpaw priiflp Baku . 

.938 

86.6 

12.3 

1.1 

19440 

Ppnda nil rP«?lHlllim Rflkll . 

.928 

87.1 

11.7 

1.2 

19260 

Pwidp nil Jfl.wo. 

.923 

87.1 

12.0 

.9 

19496 

TTaq w priinp Offnin . 

.985 

87.1 

10.4 

2.5 

18146 








In its native state, crude oil is a viscous liquid, having a dark brown color. Its 
specific gravity varies from .81 to .90, and its flash-point from 75 to 95 F. It is a 
mixture of several kinds of oil of differing boiling points, which during the refining 
process are by fractional distillation separated into groups, all the constituents of 
each of which have approximately the same boiling point. After the distillates have 
been separated there remains a heavy oil, called Masut (Ostaki) or crude oil residuum, 
which has a specific gravity of from .90 to .915 and a flash-point between 175 and 
355° F. Crude oil as well as the residuum, on account of their high viscosity and 
high vaporizing temperatures, are burned only with some difficulty in internal com¬ 
bustion engines. The gas engine liquid fuels therefore generally belong to the class 
of distillates. The refiners as yet have no standard nomenclature for the various 
distillation groups. According to Hoffer* the one in the appended table is perhaps 

the best known. , . , . TT ^ „ 

According to government regulations in Germany and Austria-Hungary,^ on y 

such of the distillates as show a flash-point of at least 70° F. when tested in the 


1 From Veit, Das Erdol, p. 432. 

1 Hofer, Das Erdol and seine Verwendung. 

































528 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


Abel instrument and referred to 760 mm. barometer, 
•can be used as kerosene oil for lighting purposes. By 
Jlash-point is meant the temperature at which the 
oil commences to give off inflammable vapor. If these 
vapors are ignited, they will burn for short intervals 
at the surface of the oil, without igniting the body 
of oil. Only when the burning-point is reached, are 
the vapors generated at a rate fast enough to main¬ 
tain a flame over the surface, i.e., the liquid itself 
burns. It follows from this that the burning-point 
is higher than the flash-point (by from 9 to 45°.). 

The heating value of crude oil and its distillates 
varies from 17 000 to 21 600 B.T.U. per lb., the 
average being about 18 900 B.T.U. When the chem¬ 
ical composition of an oil is definitely known, the 
heating value may be computed by the combination 
formula, p. 591, Appendix. This formula, using the 
accurate figures for the heating values of C and H, 
may be written 


H M = 


14 544 C+51840 



B.T.U. per lb.,. ( 6 ) 


in which C and H are expressed in parts of a pound. 

Absolutely pure lighting oil is a hydrocarbon 
having the formula Ci 3 H 2 8 , a specific gravity of 
7 - = .793. One pound of this oil therefore contains, 
using 12 as the atomic weight for C and 1 for H, 


and 


12X13. 
12X13 + 1X28 

1X28 

12X13 + 1X28 


= .847 lbs. C, 


= .152 lbs. H. 


From this the low r er heating value is 


H u -14 554 x .848 + 51840 X. 152 = 20 232 B.T.U. per lb- 


Whenever possible the heating value of such oils 
as will readily burn in suitable burners should be 
determined by calorimeter. 

The amount of air theoretically required for 
combustion is usually computed by means of the 
second combination formula, p. 588, Appendix. 
Assuming that the average composition of petroleum 
is .85 0, .14 H, and .010, this formula gives for the 
minimum amount of air required by 1 lb. of oil 


L = 


2.667 X.85+ 8 X.14 —.01 
.23 


= 14.68 lbs., 


or 


77_ 2.66 7 X .85 + 8 X .14 — .01 

^23 X .0807 


= 182 cu.ft. 


cc !2 
W + 
H U5 
< 

h3 + 
►3 .3 


i-3 

o §8 

Q 

p 00 
P3 if 
o L 

t-H 

0 i 

£ x> 

O Oh 
m c3 
H + 
<1 

§ -s 

S 0 

3 o 

o c 
« 

E 

o 


33 3 
1-0 • 
co 00 os J2 
P o 


w 

to 


s 


« 8 
® O , 

!C §0 


l© <N CO 
O <N — o < 

00 0O GO 00 GO t 


A , 


H a> 


13 s 

GO 

1 <N O 
1 00 00 


S3 


5 

O -P 

o 

03 ‘ JS 

W.g * 
_r o c 

i w 

"o cj M 

P .S 

5, CO 


3 

ci 

& 

£ 




1 3 

(3 


8_r - S « 

" 03 + O B- 

ci o o a S 
W « -g 


3 

t 

1 is 

>3 to 
33 + - 


adco^oc 
i, N © CG o 


I si 
0 $ * 
woo 


c .3 

a a 

; A » 

§ ^ 

xe> 


1 + 


NNN 
CO 00 © <N CO 
CO CD t- f-. 


t». (M 
‘OONON 

co co co t— 


-g Ci Oh 
+ -g ci 
OhA «? 

| 

c 3 a> 

3 II 

"3 _o o 

5 g + 

a . 


0 


Oh_ 
'05 'o 
3 ® w hfi 

3 c .3 .3 .3 
g '3 "q 'p *c 
■is c a h, 3 

o o ci .2. 

Ph pq o h3 o 

































FUELS 


529 


Since the chemical composition of all crude oils and distillates varies only within 
very narrow limits, it may be assumed for all of them with sufficient accuracy that 
the minimum air required for their complete combustion is 15 lbs. per lb. or 185 cu ft 
per lb. 

In determining the ratio of oil to air, it should be remembered that the former 
before or during the formation of the mixture changes from the state of a liquid to that 
of a vapor, and that its volume consequently increases considerably. The volume of 
vapor resulting from a given weight of oil increases with the temperature of vaporiza¬ 
tion. For any one temperature the lighter the oil, that is the more volatile constit¬ 
uents it contains, the greater the volume of vapor. To test the vaporizing qualities 
of an oil, fractional distillation is resorted to, during which the quantity of vapor 
evolved between certain temperature ranges is determined. Table 133 contains some 
detailed information on the vaporizing properties of certain crude oils. Accurate deter¬ 
minations of the volume of oil vapor per pound of oil are rare. One of the older 
numbers of the Gas-Journal (1895, p. 22) gives the data contained in Table 134, 
which originally seems to have referred to the manufacture of oil gas. Since the 
vaporization of oil in engine cylinders can hardly be as perfect as in special gas 
retorts, the figures of Table 134 will necessarily have to be reduced somewhat to 
apply to gas-engine practice. 


Table 133 


DISTILLATION TESTS ON MINERAL OILS 



^ Specific 

Gravity, 

° Beginning of 
. Vaporization, 

Volume Per Cent Vaporized. 


Up to 

302° F. 

from 

302 

to 

392° F. 

from 

392 

to 

482° F. 

from 

482 

to 

572° F. 

from 

572 

to 

608° F. 

above 

608° F. 

Alsatian “ Brilliant” petroleum 

“Kaiser” oil. 

Pennsylvania kerosene. 

.801 

.795 

.800 

296 

212 

212 

.8 

15.8 

30.5 

29.7 

22.0 

44.7 

32.3 

19.25 

20.2 

26.3 

16.8 

3.8 

11.7 

26. 

1 

15 

| Tests by C. Engler and 

C. Schestopal 

Bavarian crude oil. 

Roumanian illuminating oil . . 

Galician solar oil. 

Hungarian "blue” oil. 

German “red” oil. 

German yellow paraffin oil ... . 

German solar oil. 

German benzol. 

.827 

.815 

.874 

.836 

.870 

.860 

.825 

.873 

Up to 

14.7 
28.2 

20.1 

13.8 
212° F 

12.1 

15.2 

15.0 
3.7 
2.4 
57.4 
. 68%, 

10.7 

12.8 
2.7 

12.6 

38.9 

55.0 

25.5 

from 2 

9.1 

24.3 

65.1 

12.2 

39.3 
37.0 

1.0 

12 to 30 

7.8 

7.7 
14.0 

4.3 

12.3 

2.8 
.1 

2° F. 2i 

5.4 

2.4 

3.9 % 

H u = 18250 B.T.U. 
Hu = 18320 B.T.U. 
Hw = 17750 B.T.U. 

H u = 17470 B.T.U. 

Tests made by Ma- 
schinen-fabrik. 
Augsburg 


Table 134 


QUANTITY OF VAPOR OBTAINED FROM MINERAL OILS 



B-Petrol- 

Naphtha, 

7 =.730. 

Lighting 

Oil, 

r=.so7. 

Heavy Oil, 

r=w. 

Heavy Oil, 
r = .884. 

Vaporizer temperature, ° F. 

1112 

1562 

1112 

1562 

1112 

1472 

2012 

932 

1112 

1562 

1 cu.in. of oil furnishes cubic inches of vapor . 

451 

625 

469 

582 

401 

513 

594 

213 

368 

675 

1 lb. of oil furnishes cubic feet of vapor. 

10.4 

13.8 

9.3 

11.0 

7.6 

9.7 

11.3 

3.7 

6.7 

12.3 

Residue not vaporised, %. 

11.4 

5.1 

21.4 

7.5 

28.5 

12.2 

18.0 

62.3 

41.5 

9.4 

























































530 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


In-approximate calculations, it may be assumed, for light and medium heavy oils, 
that the volume increase on vaporization is in the ratio of 400 to 1 at vaporizer 
temperatures, while for ordinary room temperature the ratio may be taken at approxi¬ 
mately 200 to 1. The theoretical minimum ratio by volume of oil to air is accord¬ 
ingly from 1:30 to 1:25. Heavy oils require a vaporizer temperature exceeding 1470° 
F., as Table 134 shows, in order to obtain fairly complete vaporization. This explains 
in a measure the difficulties encountered in attempting to use these oils in gas 
engines. 

2. Alcohol. Alcohol is the youngest and as yet the least important of our gas 
engine fuels. It is- produced by fermentation of grape sugar, or the latter’s raw 
material, like potatoes or grain. Theoretically 100 parts by weight of grape sugar 
should yield 51 of alcohol, but the actual yield is from one-fifth to one-third less 
than this. Pure alcohol, absolutely free from water, called absolute alcohol, has a 
specific gravity of .7946, a boiling point of 172.4° F. and a molecular weight of 46. 
The chemical composition may be written C 2 H 5 + 0H=C2H60. One pound of C2H 6 0, 
with the usual assumptions for the atomic weight of C and H, then contains. 

19 y O 1 vfi IfiYl 

-^- = .522 lb. C, —.130 lb. H, and ±^p = .348 lb. O. 

The air theoretically required will be, by means of the combination formula, 


- 7 ^27661"X"522-+-8X. 130-.348 
.23 


9 lbs. per lb. or 


~ 112.3 cu.ft. per lb. 


The heating value of alcohol can not be accurately computed from its chemical 
composition, because the molecular grouping of the atoms C 2 H 6 0 is not definitely 
known,. Direct measurement by calorimeter is therefore the only satisfactory way. 
According to the experiments made by Thomson, the higher heating value of alcohol 
vapor, having a density of 1.601 (air=1.0), was 13 320 B.T.U. per lb. By subtracting 
from this value the heat of vaporization of the water of combustion, we find the 
lower heating value equal to 13320-1.17x1090=12045 B.T.U. per lb. Prof. E. 
Meyer gives the lower heating value as H u =11 664 B.T.U. per lb., basing his compu¬ 
tation upon the work of Favre & Silberman, who give the higher heating value at 
H 0 =12 931 B.T.U. per lb. The correction for water of combustion is taken at 
1.174X1079 = 1267 B.T.U. Meyer’s figure gives as the lower heating value of 1 gallon 
of absolute alcohol 77 089 B.T.U. 

Commercial alcohol always carries a certain quantity of water, and has therefore 
a greater specific gravity and a lower heating value than absolute alcohol. The 
degree of dilution, that is the percentage of water carried, is to-day in most cases 
found by means of a hydrometer, which determines the per cent by weight of water 
present on the basis of specific gravity, rather than the. _per_ cent by volume. 1 The 
standard temperature for weight per-cent determinations is 15° C.; for the volume- 
per cent measurement, 60° F. = 15|° C. If the hydrometer is used at any other 
temperature than T5° C., the reduction to the standard' may be made by use of 
Table 135. 


1 In the United States, the. percentage of water is usually stated in volume-per cent, thus 90% alcohol 
means that the alcohol contains 10% by volume of water. 








Table 135 

OBSERVED READING OF WEIGHT HYDROMETER 


FUELS 


531 


©0 


04 

C5 


o: 


© 

cs 


© 


X 

X 


X 


CD 

X 


iO 

X 


X 


©0 

X 


*1 

X 


X 


o 

X 


© 


X 


!>• 

!>• 


CD 

I - 


»o 

!>• 


I'* 


?© 

L>« 


©1 

I'. 




o 

o 

LO 


o 

Q 

K 

O 

P 

Q 

H 


£ 

H 

O 

rr? 


Eh 

K 

O 
►—« 
a 
fe: 


p 

o 

M 

O 

Q 

P 

<*5 

O 

S3 

Eh 

0 

£ 

w 

rrf 

—H 

Eh 

X 

w 

p 


— 


go 


o 


CD 



X 

CD 

04 

X 


04 

CD 


X 


04 

CD 

rt* 


X 


X 


04 

X 

CD 

04 

X 

X 

X 

X 



td 


CD 

CD 

CD 

CD 

LO 

d 

LO 

•o 

d 

d 


X 

X 

'V*' 

04 

04 

04 


*— 8 


d 

d 

o 

05 

C5 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

X 

X 


04 

X 

CD 

i 

X 

CD 

04 


CD 



X 

CD 

04 


CD 



CD 

rt< 


X 

TtH 

04 

X 


04 

X 

rd 

rd 

td 

CD 

CD 

CD 

CD 

GO 

iO 

d 

to 

d 


-v 

X 

X 

X 

X 

04 

04 

04 

•-H 

i-H 

—H 

o 

o 

o 

05 

05 

05 

X 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

X 

X 

X 

X 

X 


04 


X 

TjH 

04 

X 

CD 

04 


X 


04 

X 

CD 

04 


CD 



CD 



X 



X 


04 

X 

f ^ 


CD 

CD 


d 

d 



d 


X 

X 

X 

04 

04 

04 

04 


*-H 

•-H 

O 

o 

o 

05 

05 

05 

X 

X 

X 


05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

X 

X 

X 

X 

X 

X 

X 

X 

CD 

Tt< 


X 

CD 

04 


CD 

tH 


X 


04 

X 

CD 

04 


CD 



CD 

r}H 


X 



X 



X 

16 

d 

d 

‘O 



d 

-f 

X 

X 

X 

04 

04 

04 


»“H 

*— H 

>-H 

o 

d 

o 

05 

05 

05 

X 

X 

X 

d 



CD 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 


X 


04 

X 

CD 

04 


X 


04 

X 

CD 

04 


CD 

04 


CD 



CD 



CD 



X 

TjH 


CD 

iO 


-f 

-F 

X 

X 

X 

X 

04 

04 

04 

» < 

«■— 4 

r-H 

•-H 

o 

o 

o 

05 

05 

05 

X 

GO 

X 




CD 

CD 

•— 

d 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 


X 

CD 

04 


CD 

Tt< 


X 

TfH 

04 

X 

CD 

04 


CD 



CD 



CD 



CD 



CD 



CD 

-f 

r^*i 

“V^ 

X 

X 

04 

04 

04 

»-H 

*-H 

*-H 

o 

d 

O 

o 

05 

05 

05 

X 

X 

X 

d 

d 


CD 

CD 

CD 

»o 

LO 

LO 

d 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

04 

X 

CD 



X 

Tf 

04 

X 

CD 

04 

X 

CD 

04 


o 

< Tt< 


CD 

T}< 


CD 



CD 



CD 



CD 

X 

04 

04 

04 

04 

r—H 


T“1 

o 

d 

o 

05 

05 

05 

05 

GO 

cc 

X 




CD 

CD 

CD 

d 

LO 

LO 

d 

d 


CO 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

05 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

04 


CD 



X 

Tt< 

04 

X 

CD 

04 


CD 

04 


CD 

■'O 1 


CD 



CD 



CD 

rf 


CD 

Tt< 


CD 

04 

04 




o 

o 

o 

05 

05 

05 

05 

X 

X 

X 

d 

d 


CD 

CD 

CD 

d 

d 

LO 


rfH 


X 

X 


04 

w* 


05 

05 

05 

05 

05 

05 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

Tt< 


X 


04 

X 

CD 

04 

X 

CD 

04 


CD 

Tt< 


CD 



CD 



CD 



CD 



CD 

04 


CD 



o 

o 

o 

05 

05 

05 

X 

X 

X 

X 


d 


CD 

CD 

CD 

d 

d 

LO 

d 

d 


X 

X 

X 

04 

04 

04 



05 

05 

05 

05 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 

X 



X 

rf 

04 

X 

CD 

04 


CD 

04 


CD 



CD 



CD 



CD 

rt< 


CD 



CD 

04 


CD 

o 

05 

06 

05 

X 

05 

X 

05 

X 

X 

X 

X 

X 

X 

X 

88 

X 

d 

X 

87 

CD 

X 

CD 

X 

86 

d 

X 

IfJ 

X 

85 

X 

d 

X 

84 

X 

X 

X 

X 

83 

04 

X 

04 

X 

82 

X 

X 

81 

d 

X 


04 

X 

o 

04 


CD 

04 


CD 



CD 



o 



CD 



CD 



CD 

04 


CD 

04 

X 

CD 

05 

X 

05 

X 

X 

X 

X 

X 

X 

X 

88 

X 

87. 

87 

CD 

X 

CD 

X 

86 

d 

X 

d 

X 

IO 

X 

d 

X 

d 

X 

tJH 

X 

X 

X 

X 

X 

83 

04 

X 

04 

X 

82 

X 

X 

X 

d 

X 

80 

05 

1^ 

05 

-t 

04 

X 

o 

04 


CD 



CD 



CD 

Ttn 


X 

rf 


CD 

T*J 


CD 



o 

04 


CD 

04 

GO 

CD 

X 

X 

GO 

X 

r- 

X 

r- 

X 

r^ 

X 

87 

d 

X 

CD 

X 

86 

ID 

X 

d 

X 

85 

d 

X 

Tt< 

X 

84 

X 

X 

X 

X 

83 

04 

X 

04 

X 

82 

X 

X 

X 

o 

X 

d 

X 

80 

05 

^5 

d 

X 

t- 

GO 

CD 

04 


CD 

Tt< 


CD 

rjn 


CD 



CD 



X 

rJH 


o 



CD 



CD 

04 


CD 

04 

X 

CD 

td 

X 

X 

87 

CD 

X 

CD 

X 

86 

d 

X 

d 

X 

85 

d 

X 

d 

X 

84 

X 

X 

X 

X 

83 

04 

X 

04 

X 

82 

’-H 

X 

X 

— 

X 

d 

X 

d 

X 

80 

05 

d 

79 

X 

u- 

OC 

1^ 

d 

1^ 

d 

I s * 


CD <N 


CD 


O tJ< 


O 


X 


X 


o o CD LO lo lo 
X X X X X X 


xxxxxxxxxx 


rt< CD rJH CD CD (N CO N 00 O 

X N N o O 

u- u- u- i> 


i-H • _ h oD? *^0 cD 05 05 05 GO GO 

xxxxxt-^t^i^t^t^ 


CO 04 


CD ^ 


X 


X ^ X 


X 




CD rt« 


CD Tf 


CD 04 


X X X X X X 


CO W ^ Ol Ol Ol H h h o 

xxxxxxxxxx 


000505© XX Xt^t^ 


CD 04 X CD 

U- CD CD d d 
r>-i> i> 


CD 


CD 


X ^ 


X 


X ^ 


X 




CD -f 


CD ^ 


CD 04 


^ CO CO X 

00 X X X X X 


010104'H' H 'HOOOp 

xxxxxxxxxt- 


03OXXXNNNOO 

i>. t— r- i> u~ i"- i^* i> 


CD 04 X CD 

c d d d d d 

U>- t-» t>* U- 


CD ^ 


X ^ 


X ^ X ^ x -t< 


X 


CD 


CD 04 


X X X 04 04 X 
X X X X X X 


HHH0000 05 05X 

xxxxxxt^r^t^t^ 


XXNNNOCDCOiOiO 


CD 04 X ^ 

»D rt< X X 

u- 1> I> 


CD ^ X ^ 

r^i 04 04 —< •—< 04 
X X X X X X 


X ^ 


X Tt< 


X ^ 


X 


CD ^ 


CD -f 


O 04 


d©0©05©XXXI> 

xxxt^t^r^.t^t^t^i> 


N h O O O LO »0 i-O i ^ 

r>. i>. r- r*- i> i> i> 


0 04 X-f 

^ X X 04 04 

1> I> l> 1- 


CD -V 


X ^ 


CO X ^ 


X 


r—4 r*H r—4 C) 

X X X X X X 


X ^ 

©©©xxxr^t^r^co 

^ ^ ^ (>• t'w t'" 


CD rf< 


CD -t 


CD 04 


cDCOlOlOlO^'-F'^COCO 
£>.*>• t> t>- !>• t> t"» t^- I''* 


CD 04 X ^ 

X 04 04 h H 
u- i> t>- i» 


CD ^ 


X tJH 


o O O C5 05 05 
X X X 


CO X x ^ x 
COXXNUN^i?^ 

I> L- L- N N N N N 


CD 


CD rjn 


O 04 


10 »O^^^XXXIM 01 


O 04 X rfi 

04 -h *-h O O 
t- 


x ^ 


X ^ 


X ^ 


X 


C5 05 05 X X X 

t '- l>- 1 >» i> o- l> 

oo X ^ 

x GO X N N 

jv, ^ t>- L- N l''- 


00 04 X Tt< 

rd t- cd -o cd >o »o o 
L>- t>- 4> t'- *> ^ 

X 

js. j>. f>» t>» N L* N N 


Tt< 


CD rtH 


CD rH 


CD 04 


CD 04 X ^ 


rf< X X X 04 04 04 i- 
£>. i>» I'* 1^- N N L 


-h O O 


o 


o ^ 


CD 04 


XX^OIOI^hhOO 

j>, (n. ^ (n. F« F» 1^- t'* 


CO 04 X O 

O O 05 GO X 
CD CD O CD 


X -t 


X -t 


X 


X ^ 04 X rfH 04 CO ^ 

i> t» i> 


CD ^ 


CD 04 


t» h N p O O 

t> 1^" 


04 04 w 1 ^0 ^0 05 05 

N N N l'* N CD CD 


CD 04 X ^ 

C5 x x 

CD CD CD CD CD 


iQ ^ M 04 H O 


+ 


C4X^^Ot-00 05 0 


HC'lX^iOONXOO 

(i—HrHr—HrHr—Ir-H 04 


H 04 X ^ ^ 
04 04 04 04 04 























































































































































532 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 

The conversion from volume-per cent (Rtl) to weight-per cent (Gtl) may be 
made by the formula 

Ga = -,.. (7) 

in which y represents the specific gravity of the sample. The heating value of an x 
weight-per cent alcohol or a y volume-per cent alcohol may be approximately expressed 

by 


H u = 11664x B.T.U. per lb., . . (8) or H„'^77089?/ B.T.U. per gallon. . . (8a) 

An accurate computation of the heating value of commercial alcohol is not 
possible for the reasons already stated for absolute alcohol. In this case the com¬ 
putation is rendered still more uncertain by the fact that commercial alcohol often 
contains denaturizing agents. The calorimeter is therefore again the only reliable 
means of determining the heating value. 

On account of the high specific cost of heat of alcohol, see Table 118, p. 513, 
the cost of operating gas engines is considerably higher with alcohol than with either 
gasoline or kerosene, other things being the same. In order to improve the heating 
value and at the same time to bring down the heat cost, alcohol is sometimes 
carbureted with from 10 to 50% of benzol, especially for operation in automobiles. 
Although this undoubtedly brings down the operating cost, the use of benzol is 
accompanied by certain undesirable features. 1 

Liquid benzol has a specific gravity of y=.866 , a heating value of H u = 17190 
B.T.U. per lb., or H u ' = 123 790 B.T.U. per gallon. From this, for a mixture consisting 
of a% of alcohol and 6% of benzol, the approximate heating value will be 

H M = [(aXll664) +(6X17190)] B.T.U. per lb., 
or H M '- [(a X 77089)+ (6X123 790)] B.T.U. per gallon. 

If the alcohol in the mixture is not absolute, but contains a certain per cent of 
water, the corrected heating value must be used in the above equations. 

Mixtures of alcohol and benzol have not been tried to any extent in either 
stationary engines or locomobiles. The tendency here has rather been to make the 
alcohol engine as efficient as the gasoline or the kerosene engine by the use of 
higher compression. The water generally carried by commercial alcohol makes it 
possible to employ much higher compression pressures than it is feasible to use with 
crude oil or its distillates, and the efficiency of the combustion is correspondingly 
increased. It is possible in this way to largely overbalance the greater heat cost of 
alcohol by higher thermal efficiencies. 

It has been observed that the interior surfaces of alcohol engines are in some cases 
strongly attacked by rust after a short period of operation. The burned gases must 
therefore contain acid constituents. These may, during the first oxidation periods, 
consist of aldehyde or acetic acid, forming when alcohol is burned with very little 
or no excess of air. With an ample air supply even the formation of acetic acid 


1 See Authors’ Report in the “Motorwagen,” p. 271. 





FUELS 


533 


no longer takes place, the alcohol being directly oxidized to water and carbon 
dioxide, neither of which have any harmful effect on iron. 

This serves to show that alcohol engines should be built with a view to using as 
large an excess of air as possible. 


III. Fuel Mixtures 

The gas engine fuels, in preparing them for combustion in the internal combustion 
engines, are mixed with an amount of atmospheric air which makes the oxygen 
present much in excess of what is required in theory. 

Pure dry air is a mechanical mixture containing 


23.58 weight-per cent J and f 76.42 weight-per cent 1 q{ 

21.33 volume-per centI ' 1 78.67 volume-per cent I 


together with traces of water vapor and carbon dioxide. It therefore requires 


and 


4.241 parts by weight of air to obtain 1 part by weight of oxygen 
4.688 parts by volume of air to obtain 1 part by volume of oxygen. 


The weight of air is .0806 lb. per cu.ft. under standard conditions of 32° F. 
and 29.92" Hg (=760 mm.), or .078 lb. per cu.ft. when referred to 32° F. and the 
standard metric atmosphere of 735.5 mm. Hg. The specific volume, that is, the 
volume of 1 lb. of air consequently is 12.41 cu.ft., or 12.82 cu.ft., depending upon 
what are considered standard conditions. Under the average conditions of humidity, 
air, on account of the water vapor carried, is always lighter than above stated. 
Thus the average weight is found to be .0802 lb. per cu.ft. at 32° F. and 760 mm. 
Hg and 0777 1b. per cu.ft. at 32° F. and 735.5 mm. Hg. The corresponding 
specific volumes are then increased to 12.47 and 12.87 cu.ft. per lb respectively. 
To determine the specific volumes V 0 for any other temperatures (() and other 

barometer readings (5), Table 136 may be used. . . , 

The lower as well as the upper limits of the allowable ratio between air an ue 
are given, the former pretty definitely by the theoretical amount of air required by 
the given fuel, while the latter is subject to many influences and therefore difficult 
of exact determination. There is little doubt, however, that the field lying between 
the lower and upper explosive limits, that is, the so-called ‘ explosive range is 
considerably wider for similar air-fuel ratios in actual practice than has been found 
in laboratory tests or similar trials. The reason for this is the extension of the upper 
limit due to compressing and heating the mixture. Such tests, however, even if 
they are carried through in the usual way with volumes of mixture under ordinary 
pressures and temperatures, give a lot . of information, valuable to the engine builder 
in relation to the explosibility of the various fuel mixtures under varying:<*«“ 
dilution. The following has been abstracted from the more recent test reports 

The explosion limits recorded in Table 137 were determined by Dr. Eitner m 
the laboratory of the technical high school at Karlsruhe.' The figures given represent 


1 Journal fur Gasbeleuchtung, 1902, p. 1. 



534 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


Table 136 

SPECIFIC VOLUME F 0 (CUBIC FEET PER POUND) OF ATMOSPHERIC AIR UNDER 

CONDITIONS OF AVERAGE HUMIDITY 


b. 

Mercury Column. 






O p 

. 20 

25 

32 

40 

50 

60 

70 

80 

90 

100 

Inches. 

Millimeters. 

°C.- 

- 6.6 


3.9 

0 

+ 4.4 

1C 

.0 

15.5 

21.1 

26.6 

32.2 

31 

. 7 

31 

0 

787 

.4 

11 

.75 

11 

.87 

12 

.04 

12 

.23 

12 

.48 

12 

.72 

12 

.97 

13 

.22 

13 

.46 

13 

.70 

30 

.8 

782 

.3 

11 

.82 

11 

.95 

12 

.12 

12 

.31 

12 

.56 

12 

.78 

13 

.05 

13 

.31 

13 

.55 

13 

.79 

30 

6 

777 

.2 

11 

.90 

12 

.02 

12 

.19 

12 

.39 

12 

.64 

12 

.88 

13 

.14 

13 

.39 

13 

.63 

13 

.88 

30 

4 

772 

.2 

11 

.98 

12 

.11 

12 

.28 

12 

.48 

12 

.73 

12 

.97 

13 

.23 

13 

.49 

13 

.73 

ft 

.98 

30 

.2 

767 

.1 

12 

.06 

12 

.19 

12 

.36 

12 

.56 

12 

.81 

13 

.05 

13 

.31 

13 

.57 

13 

.82 

14 

.07 

30 

0 

762 

0 

12 

.14 

12 

.27 

12 

.44 

12 

.64 

12 

.90 

13 

.15 

13 

.41 

13 

.67 

13 

.91 

14 

.17 

29 

.8 

756 

9 

12 

.22 

12 

.35 

12 

.53 

12 

.72 

12 

98 

13 

.23 

13 

.49 

13 

.75 

14 

.00 

14 

25 

29 

6 

751 

8 

12 

.31 

12 

.43 

12 

.61 

12 

.81 

13 

07 

13 

.32 

13 

.59 

13 

.85 

14 

.10 

14 

35 

29 

4 

746 

8 

12 

39 

12 

.52 

12 

.70 

12 

90 

13 

.16 

13 

41 

13 

68 

13 

94 

14 

.20 

14 

45 

29 

2 

741 

7 

12 

48 

12 

60 

12 

.78 

12 

98 

13 

25 

13 

.51 

13 

.77 

14 

03 

14 

.29 

14 

55 

29 

0 

736 

6 

12 

57 

12 

69 

12 

88 

13 

07 

13 

34 

13 

60 

13 

86 

14 

13 

14 

39 

14 

65 

28 

8 

731 

5 

12 

65 

12 

78 

12 

96 

13 

16 

13 

43 

13 

70 

13 

96 

14 

23 

14 

49 

14 

75 

28 

6 

726 

4 

12 

74 

12 

88 

13 

05 

13 

26 

13 

53 

13 

80 

14 

07 

14 

34 

14 

59 

14 

86 

28 

4 

721 

4 

12 

83 

12 

96 

13 

.14 

13 

35 

13 

63 

13 

90 

14 

17 

14 

44 

14 

70 

14 

97 

28 

2 

716 

3 

12 

92 

13 

05 

13 

24 

13 

45 

13 

72 

13 

98 

14 

26 

14 

53 

14 

80 

15 

06 

28 

0 

711 

2 

13 

01 

13 

14 

13 

33 

13 

54 

13 

82 

14 

09 

14 

36 

14 

64 

14 

90 

15 

17 

27 

8 

706 

1 

13 

10 

13 

23 

13 

42 

13 

64 

13 

91 

14 

18 

14 

46 

14 

74 

15 

00 

15 

28 

27 

6 

701 

0 

13 

20 

13 

34 

13 

53 

13 

75 

14 

02 

14 

28 

14 

57 

14 

85 

15 

12 

15 

40 

27 

4 

696 

0 

13 

29 

13 

44 

13 

62 

13 

85 

14 

12 

14 

38 

14 

67 

14 

96 

15 

23 

15 

51 

27 

2 

690 

9 

13 

39 

13 

53 

13 

72 

13 

94 

14 

23 

14 

50 

14 

79 

15 

08 

15 

34 

15 

63 

27 

0 

685 

8 

13 

49 

13 

63 

13 

82 

14 

04 

14 

33 

14 

60 

14 

89 

15 

18 

15 

45 

15 

73 

26 

8 

680 

7 

13 

59 

13 

73 

13 

93 

14 

14 

14 

43 

14 

71 

15 

00 

15 

29 

15 

57 

15 

85 

26 

6 

675 

6 

13 

70 

13 

84 

14 

04 

14 

25 

14 

54 

14 

83 

15 

11 

15 

40 

15 

69 

15 

97 

26 

4 

670 

6 

13 

81 

13 

94 

14 

15 

14 

36 

14 

66 

14 

94 

15 

23 

15 

52 

15 

81 

16 

10 

26 

2 

665 

5 

13 

90 

14 

05 

14 

25 

14 

57 

14 

76 

15 

04 

15 

34 

15 

64 

15 

93 

16 

22 


Table 137 

EXPLOSIVE RANGES OF PURE AIR-FUEL MIXTURE 



Carbon 

Monoxide. 

Hydrogen. 

1 w 

Water gas. 

J Acetylene. 

Illuminating 

Gas. 

■ Ethylene. 

Alcohol, 95.14 
Weight, %. 

Methane. 

Ether. 

Benzol. 

Pentane. 

Gasoline. 

Upper explosive limit ( 1 ). 

16.5 

9.45 

12.4 

3.35 

7.9 

4.1 

3.95 

6.1 

2.75 

2.65 

2.4 

2.4 

• 

Ratio - by volume . .. 

gas 

5.1 

9.6 

7.1 

28.6 

11.7 

23.4 

24.3 

15.4 

35.4 

36.7 

40.7 

40.7 

Lower explosive limit (3) 

74.95 

66.4 

66.75 

52.3 

19.1 

14.6 

13.65 

12.8 

7.7 

6.5 

4.9 

4.9 

. air 

Ratio-by volume . .. 

gas 

.33 

.51 

.6 

.91 

4.24 

5.85 

6.33 

6.81 

11.99 

14.38 

19.41 

19.41 


























































FUELS 


535 


volume-per cent, lines 1 and 3 showing the per cent by volume of the total volume 
of the mixture represented by the fuel gas. It will be noted that the first four gases 
have by far the greatest explosive ranges, after that comes illuminating gas, whose 
19.1-7.9 11.2 1 , 

Tange is only = gg ^ = appr. — that of carbon monoxide. The narrowest 

Tange is shown by the hydrocarbons. The volume of air at the lower limits is 
considerably smaller than the theoretical amount required for combustion of the fuel 
concerned. The reason for this of course is that “ ignitability ” does not necessarily 
cease when this theoretical limit is reached. 

Several years ago a series of experiments were made at the same laboratory by 
Roszkowski under the direction of Professor Bunte to determine the effect of change 
of temperature and of contamination of mixture by carbon dioxide upon the explosion 
limits. From this extensive report 1 Table 138, showing the most important results 
bearing directly upon the subject under discussion, has been abstracted. The carbon 
dioxide-oxygen mixtures always contained, besides the fuel gas, these two gases in 
the ratio of 79 to 21, that is, in the exact proportion that nitrogen bears to oxygen 
in ordinary air. 


Table 138 

EXPLOSIVE RANGES OF CONTAMINATED FUEL GAS-AIR MIXTURES 
(Figures show the Volume-Per Cent of Fuel Gas in Mixture) 



Temp. 59° F. 

21 

2 

392 

572 

Mixture Consists of 

Upper 

Lower 

Upper 

Lower 

Upper 

Lower 

Upper 

Lower 


Limit. 

Limit. 

Limit. 

Limit. 

TTvrlrnorpn anrl ftir . 

9.5 

64.7 

9.5 

68.2 

9.6 

72.1 

9.6 

79.3 

llYUIUgCll cLAiVA an. 

TTxrrlrrtcrAn nnH — O mixture! . 

13.1 

68.1 

11.7 

69.1 

12.8 

65.1 

14.1 

61.1 

11 YvlIUtCIl dliu lW2 nnAvwivj 

Pq rlvin mnnAYiVlp qrH oir. 

14.3 

74.6 

13.2 

77.2 

12.5 

80.4 

21.0 

57.4 

Carbon monoxide and [C0 2 -0 mixture] . 

MnfViQDP orin fill* . 

21.9 

6.8 

72,5 

13.0 

20.2 

5.8 

74.8 

12.6 

26.0 

5.8 

70.0 

12.8 

38.2 

5.7 

62.9 

13.0 

"Mo+Lano onrl rCO- O mixturel. 

9.0 

11.6 

8.7 

12.0 

8.7 

12.4 

8.5 

12.2 

iVIutllallt/ dllU 2 v - / uuAviuivj. 

Tllirminafinof nrjis Qpn 9 1T . . . .. 

7.0 

22.6 

7.0 

24.7 

6.5 

26.7 

6.5 

28.6 

dllllllllLlcl Llll^ auu an. 

Illuminating gas and [C0 2 - 0 mixture.] . 

8.7 

24.2 

8.0 

26.5 

9.1 

21.7 

9.4 

18.0 


Looking at the explosion limits as determined by this table a little more in detail, 


we note the following: 

(a) Hndroqen and Air Mixture. An increase of temperature has no effect upon 
the upper limit. The lower limit, on the other hand, is extended 14.6% by heating 


from 59 to 572°. ...... ■, 

(b) Hydrogen and Carbon Dioxide-Oxygen Mixture. The explosive limit is narrowed 

more strongly by carbon dioxide than by the diluting medium nitrogen ordinary 
used. Raising the temperature has a beneficial effect only up to 212 , beyond t is 
point heating seems to narrow rather than to extend the explosive range. 


Journal fur Gasbeleuchtung, 1880, p. 491. 







































536 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


(c) Carbon Monoxide and Air. The lower limit is extended 5.8% by heating to 
392°; at 572° the range has narrowed at the lower and upper limits 6.7 and 17.2% 
respectively. 

(d) Carbon Monoxide and Carbon Dioxide-Oxygen Mixture. The explosive range 
is in this case also considerable narrower than for the carbon monoxide-air mixtures. 

The upper limit is extended 1.7% at 212°, the lower limit 2.3% at the same 
temperature, but beyond this the narrowing of the range is marked, amounting to 
4.10 and 2.5% respectively at 392°, and 16.3 and 9.6% respectively at 572°. 

(e) Methane and Air Mixture. Increase of temperature has no effect. 

(/) Methane and Carbon' Dioxide-Oxygen Mixture. The addition of carbon dioxide 
has the same harmful effect as in the previous cases. Temperature rise again without 
effect. 

(g) Illuminating Gas and Air Mixture. The temperature affects the lower limit 
only, extending it about 2% for every 180°. 

( h ) Illuminating Gas and Carbon Dioxide-Oxygen Mixture. Again the effect of 
adding carbon dioxide narrows the range both above and below. The temperature, 
raised to 572°, raises the upper limit .7% and lowers the lower limit 6.2 volume- 
per cent. 

The addition of carbon dioxide therefore has a harmful effect in all cases, and 
this conclusion naturally would hold also in case the contamination of the mixture 
is produced by the addition of burned gases remaining in the cylinder. Raising the 
temperature is of marked advantage only for hydrogen and illuminating gas mixtures, 
which advantage, however, is again lost by the addition of carbon dioxide. 

Lately Le Chatelier and Boudouard have found that very lean carbon monoxide- 
air mixtures are rendered explosive by rapid heating. While at ordinary temperature 
16% of carbon monoxide formed the upper limit, this was changed to 14.2% at 752°, 
9.3% at 912°, and 7.4% at 1112° F. 1 This determination is of importance in the 
utilization of producer and blast furnace gas (both of which are high in CO), especially 
since the explosibility of such carbon-monoxide mixtures is in itself raised by com¬ 
pression in the cylinder. 

Concerning the influence of higher pressure upon the explosive range we know 
next to nothing. Only this much is certain, that a variation of pressure has no 
effect upon some of the fuel mixtures. Thus,' for instance, Eitner and Bucerius could 
not find a change in the explosive limits of hydrogen-air mixtures by varying the 
pressure from 7 to 60 lbs., while lean carbon monoxide mixtures could be ignited 
with greater ease when compressed. 2 

The most important combustibles in producer gas proper are carbon monoxide 
and hydrogen. Both, as far as combustion in engines are concerned, have good and 
bad points, and the aim should be to balance these as far as possible in the producei 
gas itself, since in the gas-air mixture the undesirable properties of each combustible 
can only be slightly toned down. Concerning the general question as to what com¬ 
bustible gases, and how much of each, are desirable in the mixture, Kutzbach in one 
of his lectures makes the following statements: 3 

Hydrogen (H) as compared to carbon monoxide (CO) has the following characteristics: 1 
Its ignition temperature is considerably below that of CO. 2. Its velocity of flame propagation 


1 Comptes rendus, 1896, p. 1. 

2 Journal fur Gasbeleuchtung, 1901, p. 838. 

•Z. d. V. D. I., 1905, p. 238. 



FUELS 


537 


is at atmospheric pressure about thirty times that of CO. 3. Its diffusion velocity is much 
greater than that of CO, so that it mixes much more rapidly with the required air. 4. It can 
stand a much greater excess of air for combustion than CO. The presence of hydrogen in the 
mixture therefore largely determines the qualities of the combustion and also the allowable, 
although local, wall temperature in the cylinder. On the other hand, small variations in the 
hydrogen content have marked influence upon the rapidity of flame or pressure propagation 
through the charge and consequently also upon engine capacity and fuel consumption. 

From the standpoint of the engine builder, therefore, the mixture should not contain too 
much H to keep the engine from being too sensitive, nor should it contain too little, in order to help 
the somewhat sluggish combustion of CO, with its tendency to after-burning, and to accelerate the 
ignition throughout the mass. At present from 10-15% of H in the producer gas, corresponding 
to about 5% in the mixture, seems to be about the best proportion.” 

Other information bearing upon this same question may be obtained from the 
experiments made by A. Wagener upon the determination of the ratio between air 
and blast furnace gas and upon which he reported in the lecture mentioned on p. 45: 

“Of all the fuel mixtures tried that containing gas and air in the ratio of 4 to 3 showed the 
most complete combustion. This mixture burned instantaneously with the development of a 
red-blue flame. An analysis of the burned gases showed combustible remaining to the extent of 
only .7% CO. As the proportion of gas in the mixture was increased or decreased from this 
ratio, the combustion became less and less rapid, until, at the explosion limit, it burned only 
verv slowly with clear blue flame. When the ratios were gas to air = 4 :C> and 4:1.5, no ignition 
could be produced even when the pressure in the burette was raised by several millimeters of 
mercury. The blast-furnace gas used in these experiments had the following composition: 
1.5% H, 31% CO, and 9% C0 2 , the remainder being N, etc.; it was burned in open burners, that 
is, not under pressure. 

These tests also brought out the point, which, in view of Bunsen’s statements regarding 
pure carbon monoxide, can hardly seem strange, that when blast-furnace gas is to be burned in 
an open burner, such as is used in the calorimeter, the exit velocity must be kept very low. 
Experiments made later with gas from Horde on this point show that the critical velocity is 
about 2.82 ft. per sec. At this velocity the flame burns unsteadily, is lifted away here and there 
from the mouth of the burner, and now arid then goes out altogether. 

The German technical expression for mixtures consisting cf a combustible gas 
with just sufficient oxygen for its combustion is “ Knall-gas,” while if the mixture 
consists of the combustible gas with just sufficient air the term “ Luftknall-gas . is 
used. In English practice the term “true explosive mixture” has been used indiscrimi¬ 
nately for either one of these combinations. The following table, due to Professor Bunte, 1 
shows the volume relation between the combustible gases given and the volume of oxygen 
required for the true explosive mixture. 


Kind of Gas. 

Ratio Gas to 

Oxygen. 

Carbon monoxide, hydrogen, and water gas.. 

. 1 

. 1 

to 

to 

1 

2 


. 1 

to 

2\ 


. 1 

to 

3 


. 1 

to 

3 


.... 1 

to 

6 


. 1 

to 

7i 


. 1 

to 

to 

8 

1.2 




1 Journal fur Gasbeleuchtung, 1901, p. 837. 












•538 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


Multiplying the right member of each one of these ratios by 4.7, we obtain the 
theoretical volume of air for the true explosive mixture. 

Generally the excess of air in the fuel mixture should be made as large as 
possible in order to draw down combustion temperatures, prevent pre-ignition, and 
to furnish sufficient oxygen to the fuel should the mixture be imperfect. (Concerning 
the thermal advantages of lean mixture, see p. 13). The upper limit, that is the 
greatest degree of dilution, is of course set by the failure of the mixture to ignite 
explosively. In actual practice mixtures are kept well below this limit in order to 
avoid combinations that are too lean and would consequently result in low engine 
capacity. The average excess of air at normal load is usually from 30 to 40%. 
For the rich fuels even more air than this is employed, the mixtures in general having 
a heating value of from 45 to 60 B.T.U. per cu.ft. (See Part II, p. 73.) The 
variation of this heating value for various excess coefficients and for the average 
composition of the gaseous or liquid fuels discussed in the previous paragraphs is 
shown in Table 139. 


Table 139 

WEIGHT AND HEATING VALUE OF AIR-FUEL GAS MIXTURES 


The bracketed data in 
the column headings 
refer to the liquid 
fuels. 

Specific Weight in Lbs. per 
Cu.Ft. (Lbs. per Gallon.) 

Density, Air =1.0. 

Specific Volume in Cu.Ft. 
per Lb. (Gallons per Lb.) 

t- 1 

Per Cubic Foot <% 

(per gallon) £-2> 

<t> !t! 

Per Pound s? 

TO 

Air Required 
for True 
Explosive 
Mixture. 

Heating Value in B.T.U per Cubic Foot wirh 
Excess of Air Amounting to 

Cubic Feet per 
Cubic Foot. 

Pounds per 
Pound. 

of the Fuel. 

0 

25% 

50% 

75% 

100% 

125% 

150% 

B.T.U. 

B.T.U. 

Illuminating gas . . . 

.032 

.40 

31.3 

560 

17400 

5.25 

13.2 

90 

74 

63.3 

55 

48.8 

43.8 

40 

Oil gas. 

.056 

.60 

17.9 

1010 

18000 

9.50 

14.0 

96.5 

78.5 

66.5 

57.3 

50.5 

45 

40.5 

Prod, gas (coke). . . 

.075 

.93 

13.4 

135 

1800 

1.00 

1.1 

66.5 

59 

52.7 

47.7 

45 

40.5 

38.2 

Prod, gas (anthr.) . . 

.066 

.80 

15.1 

146 

2250 

1.15 

1.45 

68 

59.5 

53.3 

48.5 

45.5 

41 

38.2 

Blast-furnace gas.. . 

.078 

.98 

12.8 

101 

1320 

.70 

.72 

59.5 

53.8 

49.3 

45.5 

I 42 

39.3 

37 

Coke oven gas. 

.029 

.36 

34.5 

545 

18000 

5.00 

14.0 

91 

75 

59 

56 

49.5 

48.3 

44 

Acetylene. 

.073 

.91 

13.7 

1460 

19700 

11.9 


13.2 

113 

92 

77.5 

66.7 

59 

52.7 

47.7 

Crude oil and resid¬ 
















uum . 

7.22 


.138 

130000 

18000 


1 


70 

56 

47.7 

41 

36.5 

32.5 

29.2 

Kerosene. 

6.72 


.149 

126000 

18800 



3<ppr« 

1 K n 

73.5 

59.5 

50.5 

43.2 

38.2 

33.7 

30.3 

Gasoline. 

5.80 


.173 

109000 

18800 


I 

.u 

73.5 

59.5 

50.5 

43.2 

38.2 

33.7 

30.3 

Alcohol (abs.). 

6.67 


.150 

77700 

11660 




75 

61.2 

51.6 

45 

40 

35.5 

32 

Alcohol (90 vol. %) . 

6.90 


.145 

69000 

10000 



appr. 
q n 

64.5 

52.2 

44.3 

38.2 

33.6 

30.3 

27.5 

Alcohol (85 vol. %). 

7.00 


.143 

64700 

9250 



y .u 

60 

48.8 

41 

35.4 

31.4 

28 

25.8 















































COMBUSTION IN THE GAS ENGINE 


539 


B. COMBUSTION IN THE GAS ENGINE 

I. Theoretical Data 

For the introduction to Thermochemistry, see p. 580, Appendix. Upon the basis 
of the principles there discussed it becomes necessary first to determine the influence 
of combustion upon several of the main factors occurring in constructive and thermal 
problems relating to the gas engine. As an example we will take the average illum¬ 
inating gas for which the data is given in Table 119, p. 515, and for which the final 
results in Table 140, following, show an apparent molecular weight of m= 11.543, a 
specific heat of c p = .695, or c* = .523, also a value of a: = 1.332 and a gas constant 
R equal to 134.8. 


Table 140 

THEORETICAL THERMAL DATA RELATING TO AN AVERAGE ILLUMINATING GAS 



Composition in 

From Table 143, p. 567 in Appendix. 

V • m 

G‘Cp 

G-c v 

GR 

Volume. 
Per Cent. 

V 

Weight, 
Per Cent. 

G 

m 

c p 

Cv 

R 

H 

ch 4 

CO 

c 2 h 4 

co 2 

0 

N 

.485 

.350 

.070 

.045 

.020 

.0025 

.0275 

.0845 

.4855 

.1703 

.1093 

.0765 

.0070 

.0670 

2 

16 

28 

28 

44 

32 

28 

3.430 

.593 

.245 

.400 

.200 

.217 

.245 

2.430 

.468 

.174 

.330 

.155 

.153 

.174 

775.6 

97.25 

55.23 

55.23 

35.01 

49.79 

55.23 

.970 

5.620 

1.960 

1.262 

.880 

.080 

.771 

.2890 

.2880 

.0416 

I .0437 
.0153 
.0015 
.0164 

.2050 

.2270 

.0296 

.0360 

.0119 

.0019 

.0117 

65.53 
47.10 
9.40 
6.03 
2.67 
.35 
3.70 


11.543 
— m 

.6955 

= Cp 

.5231 
= Cy 

134.78 
= R 


The constant R may also be determined in a simpler way from 

• R 0 :R = m:mo = r : To = d:§ o, 


in which 
constant, 
hydrogen. 


50 = 775.6 mo = 2, r „ = .00559, and * = .0693 respectively represent the gas 
, e molecular weight, the weight per cubic foot, and the dens.ty (air-1.0) of 
From this we may write for the given illuminating gas, 


2 X 775.6 .00559 X 775.6 _ .0693X775 _ 134 

11.55“ “ ^0321 .40 


Again, from the difference in the specific heats, 

R= <*1^1 = 778 (.695 - .523) = 778 X. 172 ~ 134. 
A 



































540 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


The following table shows the 
as indicated in the first two lines. 

constants for 

the air-gas 

mixtures, the 

ratios being 

till* 

Ratio — by volume, 
gas 

V= 6 

8 

10 

12 

air 

Ratio —• by weight, 
gas J 

G = 15 

20 

25 

30 

Weight per cu.ft. of mixture, lbs., 

.0732 

.0748 

.0757 

.0769 

D 184 + 53.57 G 

R= <?+i 

= 58.59 

57.40 

56.66 

56.16 

.695+ .238 G 

Cp ~ G + l 

= .2667 

.2595 

.2556 

.2527 

.523+ .169 G 

G + l 

= .1914 

.1858 

.1828 

.1803 

X = C p -r- Cv 

= 1.393 

1.397 

1.399 

1.402 


It appears from this that, as the ratio of air to gas in the fuel mixture increases,, 
there is a small decrease in the value of R, c p , and c v , while x slightly increases. 

The rearrangement of the gases incident to the combustion causes a change in 
the values of m, R, c p , c v and x. Table 141 shows the resulting figures, the com¬ 
position of the burned gases being found by means of equations (18) to (20), p. 590, 
Appendix. 


Table 141 

THEORETICAL THERMAL DATA RELATING TO BURNED GASES 


Air 






Volume ratio : . . . .. _ 


6 

8 

10 

12 

Fuel gas 


r Carbon dioxide, C0 2 

.530 

.530 

.530 

.530 

Composition of burned gases volumes.... 

I Water vapor, H 2 0 

1.275 

1.275 

1.275 

1.275 

I Oxygen, O 

. 150 

.570 

1.000 

1.410 


l Nitrogen, N 

4.768 

6.348 

7.928 

9.508 

Volume of burned gas, per cubic foot of fuel gas. V r , cu.ft. 

6.732 

8.723 

10.733 

12.723 

Volume of mixture before combustion, per cubic foot of 1 T r 

c i > Vn, CU.lt. 

fuel gas ] u ’ 

7.000 

9.00 

11.00 

13.00 

Ratio: V, to V„ . 


.960 

4 

.970 

3 

.976 

2 4 

.9785 

2.15 

55.0 

Contraction. A Vn. °7n . 


Rr — R ~ 

56.3 

55.8 

55.4 


Q 

C p (based on 1 cu.ft.) 

.0204 

.0200 

.0199 

.0197 

Main constants for the burned gases . 

c„( “ “ ) 

.0149 

.0145 

.0144 

.0142 

Cp ( “ 1 lb.) 

.2787 

.2674 

.2623 

.2566 


c v ( “ “ ) 

.2035 

.1940 

.1894 

.1851 


Cp Cp 

1.370 






1.380 

1.385 

1.387 


Cy C-y 


























COMBUSTION IN THE GAS ENGINE 


541 


The computation of the specific heats of the burned gas is made by means of 
equations (11) and (11a), p. 568. The method of doing this is shown in the following 
table for the mixture whose ratio is 1 of gas to 6 of air by volume. 

The specific heats of the typical gases are given in Table 143, p. 567, based 
upon unit volume. With the aid of these we will have for exhaust gas of composition 
shown in the first column of Table 141 of the following: 


For C0 2 , C p = . 530 X .0245 =. 0130 

H : 0, C p = l .275X .0241= .0307 

O, C p = . 150X .0191= .0028 

N, C P = 4.768X .0191= .0910 

C„ = .530X .0189= .0100 
C V = 1.275X .0185= .0236 
C v = .150X .0136= .0021 
Cp = 4.768X .0136= .0648 

2(F) =6.732 cu.ft.. S(C P )=.1375 

Cp -i 3 l 6 -.0204 

p 6.732 

C P _S?-^-.02787 
r r .0732 

2 (F) =6.732 cu.ft., 2(C„) = . 1005 

C-422_ 5 -.0149 

’ 6.732 

c„-i5!i 9 -.2035 

.0732 


.0204 _ .2787 
.0149 “'.2035 


The decrease of the gas constant R to the value B, is proportional to the 
contraction during combustion. The maximum difference in this case is about 4% of 
R The specific ^eat has increased slightly, while x is slightly lower than for the 
fuel mixtures but the difference is negligible in view of the uncertainty existing with 
regard to the variation of specific heat with pressure and temperature There t>s, 
therefore no considerable error involved in the assumption made m some of the compu- 
ST of Part 1, that the values of R, * and x are the same before and after 

■combustion. 1 

II. Older Views Concerning Combustion in Gas Engines 

■ssx £"^£*”2? V'H''{HEs 

opinions, which latter were ° n ® the ^Us of experiments on actual engines, 
considerations, and in part also upon coientific standpoint, became of the 

The question, at first of " "tppamnt"that the validity of 

greatest practical importance a , de d entirely upon the possibility of 

the patents which covered Otto's claims depended entirely^ 

C d ™ and £ ****** ^the ** - — 

Viiitiu^«t oI th^ o£ engiaes ’ 

important investigations concerning particularly in t h e cases in which they 


i Corroboration of this will be found in Zeuner, 


Thermodynamik, I, P- 401 










542 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


1. Stratification of Charge and Retarded Combustion. 

Otto, and Patent No. 532: “By first introducing air and afterward the charge, such 
stratification of the charge results that the latter near the cylinder head is composed of pure, 
readily combustible mixture, while near the piston it consists practically of burned gas. This is 
done for the purpose of retarding the combustion and to decrease maximum pressure and tem¬ 
perature. The ignition flame ignites only the particles of charge in its immediate proximity; 
after that the flame is propagated to the next charge particles, but its progress is the slower 
the further these combustible particles are apart, that is, the nearer the flame approaches the 
piston face. The burning particles of the charge transmit the heat evolved to the envelope of 
air surrounding them. The resulting tendency to expansion produces an increase of pressure 
which drives the piston. Since this pressure is a result of successive combustion of individual 
charge particles, it is produced gradually. Its action therefore is not similar to that of pressure 
suddenly produced by explosion of a charge and it is consequently also not accompanied by 
the unavoidable shocks and heat losses found in the ‘explosion’ engine.” 

Slaby thoroughly agreed with Otto’s views regarding stratification anct “ slow ” 
combustion of charge: 1 


“I am confirmed in this opinion by two arguments, first, the nature of the explosion line 
and second, the height of maximum pressure immediately after ignition. Careful investigation of 
indicator diagrams from Otto engines has shown that the closest approximations to the expansion 
curves are on the average obtained by use of the (polytropic) exponent n = 1.3. It follows 
directly from this, however, that the equation of state is such as to indicate continually a 
greater influx than efflux of heat. If, on the other hand, the wavy expansion lines of the 
diagrams obtained from Hugon and Lenoir engines are replaced by average expansion lines, close 

approximations will show for the Lenoir engines n = 1.4, and for the Hugon to = 1.6. Conse¬ 

quently no influx of heat has taken place during the working stroke, and the stock of heat 
existing in the gas at the moment of maximum pressure is only partly converted into work, 
while a second and large part is transmitted to the cooling water. In both the Lenoir and 
Hugon machines, however, we have a single explosion of the entire charge followed by expan¬ 
sion. The expansion lines of the Otto diagrams on the contrary under all conditions point 

strongly to an influx of heat, and to explain this fact by assuming gradual combustion of the 

peculiarly arranged charge seems to me most plausible. 

The assumption of gradual or slow combustion, however, finds further support in another 
The maximum explosive pressures in Lenoir and Hugon engines approximate 4 atmos¬ 
pheres (60 lbs.). On the basis of the excellent investigations made by Tresca on the composition 
of the explosive mixtures used in these engines, I have computed the pressures that should 

result from complete explosion and have obtained values which differ but little from those 

actually observed. The computed pressures exceeded those actually obtained by from 1 to H 
atmospheres. The same computation made for Otto engines shows theoretical pressures of 17 
atmospheres (240 lbs.), while actual indicator diagrams only show from 9 to 10 atmospheres 
(135-150 lbs.). (Slaby in another investigation has shown that of the total heat available in the 
charge, 56% are evolved during explosion and 44% during expansion in an Otto engine, while 
the corresponding figures for the Lenoir engine are 65 and 35% respectively.) This difference 
is perhaps indubitable proof that the first ignition does not at once produce an explosion of the 
total enclosed charge.” 

To further clear this question, Professor Slaby for nearly ten years carried on a 
seiies of scientific tests, on an S H.P. Deutz engine of the older model, to determine 

the heat distribution in the cylinder. The results of this highly important investiga¬ 

tion were published by Slaby in his “ Calorimetrische Untersuchungen fiber den Kreis- 
process der Gas Maschine,” 2 from which the writer abstracts the closing sentences 

1 Verhandlungen des Ver. zur Befdrderung des Gewerbefl., 1879, p. 38. 

2 Leonhard Simion, Berlin, 1894. 



COMBUSTION IN THE GAS ENGINE 


543 


of Part IX (p. 161, Period of Ignition), in order to show how' far Slaby’s own 
original views w'ere modified by his later work. 


“Based upon the results found, the actions during the ignition period in the Otto engine 
may be described as follows: 

1. The ignition flame which is admitted to the cylinder at the inner dead center position 
of the piston first causes the combustion of that readily ignitable part of the total charge 
located in the ignition port. Indicator diagrams taken with double springs free from friction 
show' this action by a sudden increase of pressure, although the latter may be comparatively 
small in amount. 

2. As a result of this initial partial explosion there is formed a strong flame which, with 
piercing effect, strikes out into the combustion chamber proper and fires the rest of the charge 
with considerable low'er velocity of flame propagation. 

3. The velocity of propagation of the explosively formed flame depends (a) upon the average 
gas content of the charge, increasing or decreasing with the former; ( b ) upon the piston speed 
of the engine, the mechanical agitation produced increasing the velocity of propagation; and (c) 
upon the location of the combustible parts of the charge, that is, upon the degree of uniformity 
of the charge. 

4. Combustion is complete after the piston has passed over a small part of its stroke, the 
time interval being from .03 to .06 seconds. The termination of the combustion period coincides 
with the point of maximum mean temperature in the cycle. Diagrams free from friction, and 
for w'hich the ratio air to gas is about 6, clearly indicate the beginning of the expansion period 
which is then completed w'ithout further influx of heat. 

5. The maximum temperatures do not exceed 2900° F., and since dissociation of the gas 
concerned does not begin to take place except at considerably higher temperatures, it may be 
held that dissociation does not occur. 

6. Even during the combustion period the flame comes in contact with the cylinder walls 
and a part of the heat that has become available is conducted away by them. But since this 
only amounts to from 8 to 10% of the total heat, w'e may conclude, in view of the considerable 
difference in conductivity of the metal on the one hand and of the gases of combustion on the 
other, that the contact between the two is neither thorough or very rapid. The combustion 
therefore very largely takes place in the inner core of the charge, surrounded by an en\ elop 
of indifferent gases, Otto’s own theory concerning the combustion process in his gas engine, in 
which he held that stratification of the charge (non-homogeneous concentric arrangement) actually 
existed, and that this w T as mainly the cause of the excellent economic showing of the machine, 
therefore seems entirely correct. The question whether the non-uniformity of the charge consists 
in stratification from the cylinder head toward the piston, or from the center of the charge 
toward the side walls, affects the case but little and will hardly be definitely solved. The view, 
formerly also held by the writer among others, but never to his knowdedge held by Otto himself, 
that ‘after-burning’ extended into the period of expansion must, however, be considered 
erroneous.” 


The Deutz Tests. To prove the correctness of Otto’s view's, a 4 H.P. Deutz 
engine w'as so arranged that the mixture could not only be exploded by the usual 
ignition flame in the cylinder head, but also through a port which had its opening 
within the cylinder close to the piston in its inner dead center position. In these 
tests the usual ignition from the cylinder head always gave normal diagram wit 
explosion at the dead center, while with ignition from the side the diagrams always 
showed an irregular and retarded combustion, the engine horse-power decreasing >y 
one-half. If the engine was allowed to draw in first the fuel mixture and at the end 
some air, ignition from the cylinder head failed completely, while lateral ignition acted 
as before. These actions seem to justify the conclusion that the original method o 
charging proposed by Otto actually results in a charge which grows leaner toward the 


piston. 

Dewar. 

Prof. Dew r ar 


Dunns the patent litigation in England concerning the 4-cycle monopoly, 
and several other experts tried to substantiate Otto’s views by taking. 


544 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


samples of the mixture from various parts of the cylinder and examining these with 
regard to their combustion and combustibility. The result, in general, showed that the 
mixture at the place of ignition next to the cylinder head showed on the average 10% 
of combustible gas, while near the piston an average of only 5% was found, and an 
intermediate location averaged 7%. The sample from the location first mentioned 
ignited very readily, while those from the intermediate location could be ignited only 
with difficulty, and the mixtures taken near the piston failed altogether. 

Teichman somewhat later repeated these eudiometric tests on a still larger scale. 1 
The samples were as before taken at the instant of maximum compression from three 
places of the combustion space and electrically ignited in a eudiometer pipette. Of 
the resulting products of combustion the water vapor will condense to water while the 
carbon dioxide, owing to contraction, occupies a smaller volume. The total contraction 
is a relative measure of the combustible gases contained in the original sample. 
Teichman, for example, finds for the mixture taken from the explosion port (rich 

mixture) a contraction of from 12-14%, for the samples taken half way (average 
mixture) from 6-8%, and for the mixture near the piston face (lean mixture), only 

from 2-3%. Other tests consisted in taking indicator cards with varying location of 
the igniter and several forms of explosion port, and, to judge from the shape of the 
card, both the effectiveness of the ignition and the behavior of the fuel mixture. The 

final results of these tests are stated by Teichman as follows: 

“I think I have proven that the various parts of a gas engine charge introduced one after 
another will not mix uniformly, but that the charge will show a composition varying with the 
locality. Although the composition is not at all definite and cannot in any given case be 

predicted with any degree of mathematical accuracy, we do know that with proper form of 

cylinder and a certain method of charging we may expect to find with certainty in any given 
locality a mixture which will meet certain stated requirements. I further find that the local 
differences in the composition of the charge are not destroyed by compression, that they have 
a marked influence upon the entire progress of the combustion, and finally that by certain 
methods of charging, certain forms of combustion chamber and of explosion ports, the com¬ 
bustion may not only be influenced, but to a certain extent suitably controlled.” 

2. For and Against Dissociation. 

Wedding (in a discussion following the article by Slaby, mentioned p. 542): “It is difficult 

to understand how, after the strong shaking up that the entire charge must receive from the first 

explosion, there should still be local differences in composition sufficiently marked to cause the 
remainder of the combustion to proceed gradually, and also hard to see why the diffusion is not 
.sufficiently perfect to assume the combustion complete at least after a second explosion. It 
seems to me that the law of combustion established by Bunsen, according to which only so 
much gas can burn as is required to maintain a certain temperature, at which temperature dis¬ 
sociation of the products of combustion sets in and further combination (combustion) cannot 
take place, offers a sufficient explanation for the facts cited by Prof. Slaby. That which Bunsen 
has proven in the case of carbon monoxide and air, certainly also applies to illuminating gas 
and air.” 

Clerk. 2 “The cause of the sustained pressure shown by the diagrams is not slow inflamma¬ 
tion (or slow combustion as it has been called), but the dissociation of the products of com¬ 
bustion, and their gradual combination as the temperature falls, and combination becomes possible. 
This takes place in any gas engine, whether using a dilute mixture or not, whether or not 
compression is used before explosion, and indeed it takes place to a greater extent in a strong 
explosive mixture than in a weak one.” 


1 Z. d. V. D. I., 1887, p. 271. 

2 Proceedings Inst. Civil Engrs., 1881-82. 



COMBUSTION IN THE GAS ENGINE 


545 


Slaby. 1 “It is well known that all gaseous combinations break up into their constituent parts 
at a certain temperature, and that at this temperature these constituents are not again capable 
of combination. There is no doubt that dissociation plays a certain part in the action of a gas 
engine. As a result of the explosion, a strong influx of heat raises the temperature to such a 
point that some of the parts of the gaseous charge will commence to. break up; beyond this 
point combustion cannot proceed and a further increase of temperature is hence impossible. This 
process is not combined with a loss of heat, because as soon as the temperature in the engine 
falls, the dissociated particles may again unite and thus render further combustion possible. The 
result of dissociation, however, is that only part of the illuminating gas really burns at the 
moment of explosion, while a further part does not burn until after expansion has started. 
According to this, dissociation would mean that there is an influx of heat along the expansion 
line and not a withdrawal. But there are two agencies operating at the same time: the cooling 
water abstracts heat, while dissociation, or rather the disappearance of it, furnishes heat. As a 
final result, heat may either be supplied or abstracted. The nature of the expansion line will 

show whether one or the other is the case. ... 

Schottler. 2 3 “ I have never been able to thoroughly believe in any considerable effect due to 
dissociation, for if combustion really ceases at a certain temperature, it must recommence when 
the temperature sinks below this level, and presupposing uniform composition of charge, the 
expansion line must therefore necessarily be an isothermal as long as combustion is not complete. 
That is however, never the case in a gas engine. But above all. the temperatures at which, 
according to Mallard and Le Chatelier, dissociation ensues, are always much higher than the 
maximum temperatures which occur^ in a gas engine. The latter is probably always below 

2700° F. and in no case much higher.” ....... , • 

Witz 3 “ It is not necessary to fall back upon the phenomena of dissociation to explain 

prolonged’ influence of the burning part of the charge upon the burned gases. As a matter of 
fact, the condition mentioned occurs under circumstances which render dissociation impossible, 

since the temperature in the cylinder does hot exceed 2600° F ” 

Mallard and Be Chatelier. “The cooling curves obtained by us have enabled us to determine 
with certainty that when a quantity of carbon dioxide which was raised to combustion tem¬ 
perature, is allowed to cool in a closed vessel, the dissociation which occurred at the —t t 
combustion steadily decreases until the mean temperature of the gas is 1800 C. (3-/0 1.). 
That at least is what occurs when the density of the carbon dioxide corresponds to that at a 
pressure of 50 cm. Hg. at 15“ C. (19.7" Hg. at 59“ F.). The average temperature of the gas at 
5S dissociation disappears decreases with the density of the gas; .t amounts MOJ. 

(2920“ F.) when the density corresponds to that at a pressure - ■ -S w haT ' e 

1 tin ond U onlv 1160° C. (2120° F.) when the pressure is 13 cm. (5.2 ) Hg. VVe na\e 

SrHS ^ f2£s5 

the carbon dioxide had been diminished, and the mean 

appears decreases, all other conditions tlfan with carbon 

nitrogen than with the ° th (, r concer ned we have not been able to prove any considerable 

r;r„ e f - - 

in a cLd Vessel." 

3. Uniform Mixture and Rapid Combustion: for and against the use of the Explosion 

P ° rt ’ , X3 + o 7 o, “When an easily inflammable mixture only contains just 

Otto, and Patent No. 2/35. combustion in the mixture after ignition will be 

excite mixture of 

tasSatXr^ “enge^T a shorter time depending upon the amount of air added. 


* Lecture before the General Convention of Gas-men, 1883, 

2 Z. d. V. D. I., 1886, p. 253. 

3 Ann. de Chimie et de Physique, 1883, vol. 3U. 





546 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


The charge in a gas engine cylinder, not stratified but of more or less uniform composition, 
must therefore be considered a diluted mixture of this kind. The charge may yet be readily 
ignited, but the combustion proceeds at a slower rate than that found in a stratified charge, 
because the rapid propagation of the combustion due to the explosion of the rich mixture near 
the place of ignition no longer occurs. It is possible, however, to explode such a diluted uniform 
mixture with the required greater rapidity, when, instead of starting the combustion in a single 
small locality, ignition is made to take place at the same time over an extended distance or 
area. For this purpose the combustion space of the cylinder which contains the lean gas mixture is 
connected with a branch chamber (port) which contains a rich mixture. If the latter is now 
ignited, the flame produced by this explosion will shoot out of this branch chamber or port into 
the combustion chamber proper and in its path will induce ignition in the diluted charge with 
rapidity and in all directions. By varying the volume ratio of the rich mixture to the lean 
main'charge it is possible to regulate the rapidity of the combustion of the entire charge, thus 
adapting the duration of the combustion to the piston speed or controlling the pressure at begin¬ 
ning of expansion.” 

Clerk. “It does not matter whether the mixture used is rich or weak in gas; the rich 
mixture can be fired slowly and the weak one rapidly, just as may be required. The rate of 
ignition of the strongest possible mixture is so slow that the time of attaining complete 
inflammation depends only on the amount of mechanical disturbance permitted. 

Fig. 685, a diagram from an Otto engine, shows what happens when the ignition comes late 
and the movement of the piston overruns the rate of the speed of the flame. The normal lines 

are those in which the rise is almost straight up from the point of the beginning of the ignition; 

they are marked a and 5; the line c, although commencing 
from the beginning of the stroke, does not record the 
maximum pressure till the piston has moved forward one- 

third of its stroke, while the line d does not depart from 

the compression line until one-third of the forward move¬ 
ment, and does not attain its maximum till near the end 
of the stroke. In the last case the ignition has been missed 
until the piston is in rapid motion, and consequently the 
flame is at first unable to overtake it. This slow com¬ 
bustion, or rather slow ignition, in the gas engine must 
be avoided, and every effort should be made to complete 
the combuston as soon as possible after ignition. The 
perfection of slow combustion would be attained when the flame spread ust as rapidly as 
the piston moves forward, and the pressure was never raised above that due to compression. 
The pressure diagram would then give the ideal results of ‘gradual expansion of gases’ 
and a ‘perfectly sustained pressure.’ But this is just the condition of greatest loss of 
heat: sustained pressure means sustained, indeed, increasing temperature, and the object to be 
attained in a good gas engine is to produce the most rapid possible fall of temperature due to 
work performed, to keep the mean temperature as low as possible, and it is only so far as this 
is successfully done that economy is possible. Slow inflammation causes loss of heat and power; 
rapid inflammation reduces the loss to a minimum while attaining the maximum possible power. 

The modern gas engine does not use slow inflammation (or slow combustion, if the term be 
preferred), but when working as it is intended to do, completely inflames its gaseous mixture 
under compression at the beginning of the stroke. By complete inflammation is meant complete 
spread of the flame throughout the' mass—not complete burning or combustion. If by some 
fault in the engine or ignition arrangement the inflammation is a gradual one, then the maximum 
pressure is attained at the wrong end of the cylinder, and, great loss of power results.” 



Witz, in a series of experiments carried on in a free piston engine whose velocity 
could be changed at will, found that after-burning occurred for every composition 
of charge and manner of charging, but he also found that the more rapid the expan¬ 
sion of gases, i.e., the higher the piston velocity, the more rapid the combustion and 
the greater the maximum pressure. 1 From this he reaches the following conclusion: 


1 Ann. de Chimie et de Physique, 1883. 








547 


COMBUSTION IN THE GAS ENGINE 


This is a law of great importance concerning the theory of gas engines. As a matter of fact, how¬ 
ever, the marked influence of the piston speed is secondary to the effect produced by the walls, for 
if this were not the case how could rapidity of expansion affect explosion phenomena as indicated? 
This can only be due to cooling by contact with the metallic walls, which cooling, lasting a 
shorter or longer time, abstracts heat from the charge and reduces the violence of the reaction 
The rapidity of combustion is, however, not the only thing affected. The influence of the 
cooling action also reduces the diagram area developed, the work done is decreased, and the 
efficiency drops, as was shown above. Therefore, in order to utilize the greatest possible portion 
ot the heat contained in the explosion charge, it is important to expand the burned gases in the 
shortest possible time and the superficial area of the enclosing cylinder walls should be reduced 


c . ir superficial area 

as far as possible, i.e., the ratio ~ I ume of c harge should be a minimum. . . . The influence of 

the cylinder walls is Consequently the great regulator of combustion phenomena. It is capable of 
accelerating or retarding combustion. Any dilution of the charge naturally makes this effect 
more marked, because the mass of inert gas in which the explosive charge may be considered 
suspended acts, like the walls, as a cooling medium, but after-burning may also take place in an 
undiluted mixture. This conclusion, logically derived also from other tests, seem to us important; 
it both invalidates and supports the theory warmly advocated by Clerk. We agree with the 
eminent English engineer in that the combustion should not be purposely retarded; this retarda¬ 
tion is an imperfection which should not be tolerated, and Otto is in the wrong when he 
advocates it. Unfortunately, however, this retardation, this so-called after-burning cannot always 
be completely avoided, because—as held by Clerk—the combination of the dissociated compo¬ 
nents of the charge only permits of a gradual development of the heat generated by the 
combustion, while the writer holds that the effect of the cylinder walls can only be toned down 
but never completely eliminated.” 


Ernst Korting, in conjunction with Professor Frese, made a number of tests on 
a vertical 4-cycle engine of his own make, during which the positions of igniter and 
of inlet (explosion) port w r ere changed alternately between the cylinder head and the 
piston. Both the indicator cards obtained and the power capacity shown v T ere equally 
satisfactory for all of the tests, which could only have been the case with a very 
uniform mixture. In connection with these tests, Korting says: 1 

“The best efficiency is attained when the distribution of the charge is such that the fuel is 
completely burned, and, urther, such that the development of pressure is the most rapid possible. 
. . . Examining a stratified charge with reference to its combustion, we find: 

1. That there is a very rapid combustion of the localized main part of the pure charge. 
This is necessarily accompanied by the development of a very high temperature combined with 
considerable local cooling and corresponding heat losses. 

2. That, since combustion must necessarily stop in the zone of transition from one gas to 
another as soon as dilution in any one place has exceeded a certain limit, the use of a stratified 
charge must be accompanied by corresponding unavoidable losses due to incomplete combustion. 

Another factor speaking strongly against the use of stratification, is the fact that the form 
of cylinder head called for to carry out this method of. charging is with respect to cooling 
very bad. 

The occurrence of after-burning, besides be ng partly due to the influence of the cylinder 
walls, may in small part also be explained on the basis of an observation made by Mallard and 
Le Chatelier, according to which, under conditions of gradual progress of ignition, the chemical 
process of combustion at any given point lasts longer than it the combustion had been explosive. 
Finally, there is the possibility that imperfect mixing may lead to the formation of several single 
‘ tongues ’ of flame, such as are observed in the case of the combustion of illuminating gas in an 
unlimited air supply, and that this very fact causes after-burning, a thorough mixing taking 
place only during the combustion itself. In any case, no matter which of the thiec possibilities 


1 Z. d. V. D. I., 1886, p. 739, and 1887, p. 997. 





548 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 

above outlined may be the cause of after-burning, the constructive means for remedying it is the 
same, and that is, to agitate the mixture as strongly as possible before combustion. 

Whether the changing of the inlet port, in its capacity as explosion port, also has 
any effect upon the indicator diagram, remains problematical. It is known that the normal 
velocity of flame propagation, even in the most inflammable mixture, is about 6.5 ft./sec., 
while the pressure wave proceeds at the rate of 984 ft./sec. If, now, a mixture be ignited 
at the closed end of a port or passage, the combustion at the start causes a slight increase of 
pressure which drives the mixture ahead of it out into the combustion chamber with a velocity 
which is much in excess of that with which the ignition can possibly follow. Hence, in any 
case whether the explosion port be long or short, the flame reaches the combustion chamber 
end without ‘over pressure.’ and will spread from here through the charge as though the ignition 
had originally ensued at the mouth of the port.” 

Sckottler bases his opinions on stratification of charge and retarded combustion 
upon the results of extended tests on engines of various makes and types, and does 
not deduce them entirely from his observation on the Otto engine. Besides the latter, 
he also examined Korting and Benz 2-cycle engines, both of which work with charge 
pumps and in which any considerable degree of stratification can hardly be expected. 
Professor Schottler draws the following conclusions from his results: 

“Otto’s view regarding stratification is hardly tenable fundamentally. Taking into account 
the relative dimensions, the charge cannot very well be formed in any way except, as the piston 
moves outward, a current of air or of mixture flows into the cylinder, and this current, after 
passing the inlet port, must gradually expand, decreasing its velocity of flow. Simply considering 
that the ratio of inlet port to cylinder cross-section is as 1 to 40 in a 4 H.P. engine, one would 
be more apt to conclude that the conditions are much more favorable to thorough mixing than 
to the opposite. A stratification of charge would seem possible only when the cylinder is of 
such length that the incoming stream of charge is not able to overtake the piston. ... It has, 
however, been shown in tests on other engines that the assumption of stratification of the 

charge is not at all necessary to explain the slow drop of the expansion line, irrespective of the 

question whether such stratification occurs in the Otto engine or not. 

“To prove that stratification of charge does not exist, ignition in the Benz engine was 
produced in six different places in the cylinder. If the composition of the charge had been 
different in different places in any marked degree, this fact would have been shown by the 
diagrams. Nothing of the kind could be noticed, and the brake horse-power of the engine 
proved to be independent of the location of the igniter. In spite of the fact that not a trace 
of stratification appeared, the diagrams show the slow drop of the expansion line, indicating 
influx of heat during expansion and clearly pointing to strong after-burning. Entirely similar 
results were obtained with one of the older type of Korting engines in which not a trace of 
stratification as defined by Otto could be observed. It was noted, however, in the latter tests 
that the ignition, when produced immediately in front of the inlet port, was much more effective 
than that produced in any other place. This observation agrees with that made in the Deutz 
tests, where it was shown that ignition from the cylinder head was much better than the ignition 
from the side and nearer the piston. . . . Having now shown that after-burning is not the result 
of either dissociation or stratification of charge, the question remains as to why it occurs at all, 

or more correctly, why the expansion line falls much more rapidly in the older engines than in 

the new. This question has been answered by Witz on the basis of very careful tests. In his 
discussions upon this subject he comes to the conclusion that the shape of the expansion line is 
entirely controlled by the influence of the cylinder walls.” 

After a detailed consideration of the numerical results obtained by Witz in the 
tests above mentioned, Schottler finally concludes: 

‘that the process of combustion in itself is the same in the newer as in the older engines. 
The success of the -newer machines must therefore be sought in other directions, such as the 
use of compression, ignition at the dead center, which, in combination with compression, 


COMBUSTION IN THE GAS ENGINE 


549 


avoids shock, greater piston speeds, and finally the use of very lean mixtures, which has become 
possible by localizing a rich mixture in an ignition port, the action of which induces vigorous 
combustion.” 


III. Current Opinions Regarding the Process of Combustion 

The controversy concerning “stratification” and “slow combustion” has long 
since died down. Even Otto himself, in his Patent No. 2735, had already abandoned 
a good share of his original views and long before the three first claims of his original 
Patent No. 532 were declared invalid, the charging and combustion process protected 
by these claims had become practically of no importance to the owners of the patent 
themselves. To-day the fundamental principle of gas-engine operation is to have 

Pure and uniform mixture and most rapid combustion, 

for upon this depends not only the efficiency of the entire combustion process, but 
also the specific engine capacity. Where it becomes necessary to dilute the charge 
in order to decrease the combustion temperature, or to increase the compression limit, 
atmospheric air only should be used for the purpose. Air in excess not only favors 
the combustion of the charge but also protects against over-heating those parts of the 
interior surface hard to reach by water-cooling, and finally increases the relative 
stock of heat, increasing the charge weight by keeping the temperature of the incoming 
mixture down. 

The second most important rule of design calls for 


Highest possible compression. 

This, besides increasing the explosion pressure, also raises the mean effective 
pressure/ On the other hand, it draws down the terminal pressure, decreasing the 
exhaust loss, see p. 13. Further, on account of the smaller volume of the charge when 
ready for ignition, the proportional admixture of burned gases to the charge and the 
loss of heat to the cooling water will both be less. Hence the general effect of high 
compression is to improve the thermal efficiency on all sides, and none of the means 
of improving the gas engine available to the designer has proven of greater effective¬ 
ness than the use of high compression. . 

In striving after high thermal efficiencies, the compression pressure has been 
increased step by step from 30 to 200 and even to 300 lbs., and the latter figure 
probably comes pretty close to what is practically possible. The heat of compression 
on other counts so desirable, becomes a most serious drawback at the moment it 
raises the temperature sufficiently to reach the ignition temperature of the charge 
The latter differs considerably for different fuels and for different fuel-air ratios of 
the same fuel, and hence in every case there is one definite most economical and yet 
safe compression limit. In general, the greater the content of carbon, and‘ 
of hydrogen, in gases or gas-air mixtures, the more readily will they ignite Conse 
quently the vapors of the liquid fuels like kerosene and gasoline occupy ^most 
unfavorable position in this respect in the fist of fuels, and these arejofiowed, m t 
order of allowable compression, by oil-gas, water-gas and producer-gas. The highest 
compression pressures are reached with mixtures the main combustible of which is 
carbon monoxide This is particularly the case with blast-furnace gas, and this fact 
has materially helped in the economic development of the blast-furnace gas-engine, in 


550 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


which it was necessary to employ unusually high compression pressures from the 
beginning. 

Low suction temperatures and effective cooling during compression are the only 
two means available to decrease the end temperature of compression, and the effective 
utilization of these, especially the latter, form a subject upon which designers will 
expend their efforts for some time to come. Neglecting the Banki engine for the 

present, the cooling of the charge has up to the present been left entirely to the 

cylinder walls and the piston face. This is probably quite satisfactory even for high 
compressions in the smaller engines in which the superficial area is sufficiently large as 
compared with the volume of the charge. As the cylinders increase in size, however, 
the conditions grow less favorable (since the volume increases as the cube and the 
superficial area only as the square of the diameter), and a further complication results 
from the fact that the thicker walls of the larger cylinders offer a greater ‘ resistance 
to conduction, thus favoring local overheating. This is the reason why the smaller 
sizes of any type of engine can use a higher compression than the larger sizes, and 

also why the best economic efficiency of the smaller engines is not equalled in the 

larger machines. Resort is then necessarily had in such cases to special methods of 
cooling, the choice being either direct injection of water or artificial means of increasing 
the heat conducting surfaces of the combustion chamber, a scheme which was touched 
upon more in detail on p. 131. 

As a third requirement for a perfect process of combustion may be mentioned 

Proper form of combustion chamber and correct position of igniter. 

The constructive means which may be used to meet this requirement are illus¬ 
trated and discussed in Part II, pages 131 and 251, while p. 559 gives some figures 
based on tests. 

The degree of imperfection of the combustion in gas engines may be numerically 
determined by accurately finding the content of combustible gases in the exhaust 
gases. An investigation of this kind is complicated and hard to carry through because 
the apparatus generally used in ordinary technical chemistry either gives results 
altogether too approximate or fails completely. 1 This is the reason why any definite 
information of general interest regarding the part that the heat loss due to incomplete 
combustion plays in the general heat balance, is only now and then obtained in tests 
very carefully prepared for the purpose. The figures available, however, admit of no 
doubt that in our present day engines the losses due to incomplete combustion are 
still very considerable, and certainly higher than is generally assumed. Table 142, 
compiled from tests made by Prof. E. Meyer, 2 may serve to prove this statement. 

The remarkable thing in this table is the increasing heat loss with decreasing 
load. This, however, was in this case shown to be due to imperfect speed regulation 
(by means of taper cam, see p. 241, Part II). In general, the results show that, 
operating with illuminating gas at normal load (about 10 B.H.P.), the heat loss due 
to incomplete combustion amounts to from 3-5%, and at half load (about 6 B.H.P.) 
it is as high as 15%. The analysis of the exhaust gases made when operating on 


1 See Z. d. V. D. I., 1902, p. 948. 

2 Untersuchungen am Gas-motor, Z. d. V. D. I., 1902, p. 1303; also Mitteilungen iiber Forschungsar- 



COMBUSTION IN THE GAS ENGINE 


551 


producer gas shows nowhere near as con¬ 
sistent results on account of the very 
dissimilar compositions of the fuel mix¬ 
tures. At about 8 B.H.P. more than 13% 
of the heat in the fuel was shown in 
unbumed particles in the exhaust gases; at 
about 7 B.H.P. one of the gas samples 
showed even 48% (supposing there is no error, 
as Meyer himself remarked). In general the 
heat losses due to this cause are greater 
when operating with producer gas than 
when working with illuminating gas. This 
is what may be expected in any case, both 
because of the varying gas-making condi¬ 
tions of the generator and the comparatively 
larger volume of fuel gas to be mixed 
with air each cycle. Improvement of mixing 
and ignition processes is therefore of special 
importance in the case of power-gas en¬ 
gines. 

Ernst' Korting at one time compared 
the various events of the cycle of a gas 
engine with those of the steam engine in 
order to facilitate forming an opinion of 
the former. 1 To the writer this compari¬ 
son is so appropriate and generally valid 
even to-day, that a few of the points 
discussed will be inserted here. 


“ Every engineer knows that the best efficiency 
in a steam engine is attained when the steam 
is admitted suddenly under the highest possible 
pressure, that is, when the diagram shows a 
vertical line in the dead center position of the 
piston, and when, after the required quantity of 
steam is admitted, it is suddenly cut off, so that 
any further supply of steam, i.e., heat supply, 
to the cylinder is completely cut off during the 
expansion period. All throttling of the steam 
during admission and all supply after cut-oft 
simply decreases the efficiency. 

Every experienced engineer knows further 
that the high initial pressures, which in the 
steam engine occur at every dead center position 
very suddenly and with greater rapidity than 
in gas engines, are not, in a good engine, 
accompanied by jar or shock. The question o 
avoiding shock can, or rather need, never e con 
sidered as a reason for so controlling the process 
of combustion that the maximum pressure does 
not occur at the dead center position m gas 



i Z. d. V. D. I., 1886, p. 757. 











































552 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


engines any more than in steam engines. The means employed to avoid shock under sudden 
pressure reversal in steam engines are well known, and it is sufficient to remark that exactly the 
same means are applicable to gas engines. In the latter, however, the solution of this problem 
is much simplified by the fact that the generation of pressure in the mixture is gradual and 
never very sudden. Accurate measurements on the rapidity of combustion in fuel mixtures of 
varying strengths and under varying degrees of compression are unfortunately lacking. In any 

case the velocity of flame propagation is so low that 
the diagram, Fig. 686, taken from the combustion of a 
charge that consisted of a rich mixture of illuminating gas 
and air, shows a combustion line distinctly inclined, in 
spite of the fact that the piston velocity near the dead 
center is low. The lines of this diagram clearly show the 
difference in the time interval required for the development 
of pressure, on the one hand between a highly explosive 
mixture, consisting only of illuminating gas and air, and, 
on the other hand, between a less inflammable mixture 
made up of two-thirds air and illuminating gas and one-third 
burned gas. 

The application of the theory of construction of the steam engine to the gas engine 
therefore teaches us 

1. That, in order to attain the best efficiency, the pressure development should be made 
as rapid as possible, the heat should all be developed in the dead center position, g,nd during 
expansion no further heat should be supplied. So-called after-burning of the charge should, 
according to this, be avoided, and the writer would compare its effect to that of a leaky steam 
valve. 

2. The charge should previously be highly compressed, because the useful effect of this 
increases in a greater ratio than the work required for compressing. 

3. External cooling should be avoided. 

The third rule can, unfortunately, not be fulfilled in gas engines. Instead of protecting the 
cylinders against loss of heat it becomes necessary for practical reasons to cool them and to cool 
them so strongly that the temperature difference between ignited charge and cylinder walls is 
considerably over 1800° F. The impossibility of avoiding cooling is the main reason for the 
principal differences in the construction of gas and of steam engines. A consideration of the 
circumstances thus controlling the design of gas engines brings out the following: 

(a) While in a steam engine the highest possible steam temperature may be employed 
without unfavorably affecting the economy, a temperature increase in the gas engine is accom¬ 
panied by heavy efficiency losses due to cooling. It therefore becomes necessary to keep the 
initial temperature of the fuel mixture down. Opposed to this is the fact that a low temperature 
is accompanied by low absolute efficiency, and this in turn means low economy. It should 
further be pointed out that a low temperature of the ignited charge can only be obtained through 
mixing with large quantities of inert gases, and that in a charge so diluted the pressure develop¬ 
ment must be so slow that the fulfilment of rule 1 above is out of question. The designer is 
called upon to steer a middle course between these too conflicting conditions, he should not go 
to the limit on either side. 

(b) Since the amount of cooling is a function of time, it is one of the requirements for 
best efficiency of combustion to shorten the time during which cooling acts, i.e., the rotative 
speed of the engine should be as high as possible, a condition which is entirely without influence 
upon the efficiency of steam engines. 1 This requirement leads to the use of rich fuel mixtures in 
order to have the point of maximum pressure somewhere near the dead center position in spite 
of high piston speed. 

(c) Since the amount of cooling is also a function of area exposed, a further requirement 
for best efficiency is to give such shape to the cylinder that the ratio of superficial area to 
volume enclosed shall be a minimum. This is again a point which is without influence upon 
steam-engine efficiency.” 



In the translator s opinion both this and the following point are intimately connected with the 
question of cylinder condensation in steam engines, and therefore have some effect upon steam-engine 
efficiencies. 







COMBUSTION IN THE GAS ENGINE 


553 


Although the velocity of flame propagation in gas engine fuel mixtures differs 
considerably according to composition and other properties, and although little is 
known regarding this .question, the fundamental relations, pointed out by Korting 
in (5) and (c) above, that exist between inflammability of charge, size of the com¬ 
bustion chamber, and rotative speed of engine, may to a certain extent be fixed 
mathematically. 

Suppose that, with a certain velocity of flame propagation v, an ignition distance 
l (equal to the greatest length of the combustion chamber), and an engine speed of 
n revolutions per minute, the maximum combustion pressure occurs at the point a 
in Fig. 687, at which the combustion is complete. Then the time of combustion is 


t = — seconds, 
v 


( 1 ) 


and the angle passed through by the crank in this time is 

a , 360 360 l , 

P = tn 60 =n "60 7 degrees ' 


( 2 ) 



Assuming that the length of the connecting 
position will be 



rod is infinite, the corresponding piston 
s /?). 


If now the rotative speed n be doubled, everything else remaining the same the 
crank angle passed through to the point of complete combustion and the new piston posrtion 

will be expressed by 

360 l j. .(3) 


“=^60 v degrees> 


md 


y = r(l —cos a). 


(4) 


rhe maximum explosion pressure then occurs at the point 6 and the ignition angle 
s hence doubled along with the rotative speed. 

Area abl expresses the work lost on account of excessive piston speed. Equation 
(2) shows two ways m which this loss may be avoided, either by increasing a or by 












554 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


decreasing l. Compressing the charge works along both of these lines at the same 
time since it decreases the volume of the charge when ready for ignition (and 
consequently also l) and increases the inflammability of the compressed charge, (and 
hence also v). The higher the compression of the charge, therefore, the greater may 

be the piston speed or the leaner may the fuel mixture be made. 

A further means of producing the maximum explosion pressure near the dead 
center, mostly used with the leaner gases and also in automobile engines, consists 

in advancing the point of ignition. If any given mixture, difficult of ignition, requires 

for its complete combustion a time interval equivalent to the crank angle a, Fig. 688, 
the maximum explosion pressure will evidently occur at the point a . If, however, 
instead of causing the ignition at the dead center, the charge is fired before the end 
of the stroke, while the crank is, say, in the position /?, the combustion, proceeding 
with unaffected velocity, will then be complete after the crank has passed through 
the angle «o above center and hence the point of maximum pressure has been moved 
in from a to b, nearer the dead center. Advancing the ignition has, therefore, resulted 
in a saving of work equivalent to an area equal to [abc— cde — the area above the b — 
curve], without reference to any of the indirect effects of early ignition. This gam 
would only be wiped out when the areas of both diagrams become the same, but 
practical difficulties (shocks to the engine) put a stop to the extent to which igni¬ 
tion can be advanced. In any case this discussion shows that the proper time of 
ignition depends not only upon the properties of the charge, but also upon the size 
or shape of the combustion chamber and the crank velocity. 

But it is not only the relatively low velocity of flame propagation which 
confines the allowable piston speed to within certain limits, there is also a second 
process, that of diffusion of the various gases making up the charge, which draws 
just as definite a limit. Every diffusion of one gas in another requires a certain 
length of time, and the time available for the piocess in the case of fuel mixtures 
is by no means as extended as is commonly assumed. The known diffusion coefficients 
(see p. 583, Appendix) are not of very great service in the solution of the question, 
because they throughout refer to kinds of gas as well as to pressure- and temperature- 
conditions not found in the cylinder of gas engines. But even these laboratory 
results show that the time of diffusion is of importance also to the gas-engine 
designer, and that he should take care at least not to cut the available time down 
to less than that of one stroke (by incorrect methods of governing, for example, 
see p. 241). The later investigations of Petreano, for instance, have shown that one 
volume of methane requires about six seconds to completely diffuse in one volume of air, 
while from ten to twelve seconds are required if the same volume of methane is 
injected into five volumes of air. The theoretical volume of air required for the 
combustion of methane is, however, twice the volume of air last mentioned. It is, 
of course, true that the heating up of the charge during the suction and charging 
strokes promotes diffusion, but the advantage is again partly counterbalanced by 
the fact that the increased pressure of the compression stroke has the opposite effect. 1 


1 The mobility of the gas molecules varies inversely as the density d. For two gases having the same 

molecular weight and equal internal friction the velocity of diffusion is in general D-l. The same 

d 


law is expressed in a different form by the equation of Clausius for the so-called 

m which a is the original mean distance between the molecules (that is, a function 
distance between the two molecules at the moment of impact. 


mean travel,” L=~. —5, 
of S) and s is the center 



COMBUSTION IN THE GAS ENGINE 


555 


In any case it seems that even in slow-running engines (n = 150 to 180) the time 
interval represented by two strokes is hardly sufficient, for complete diffusion, and that 
the latter extends into the expansion stroke. This fact alone would explain after¬ 
burning. In high speed engines this lack of uniformity in the mixing must of course 
be more marked, a condition which is proven by experience. 

Considerations of this kind seem to make it desirable not to form the mixture 
in the working cylinder itself, but to previously mix the charge in a sufficiently large 
chamber unaffected by the movements of the piston and to draw the charge from 
this chamber as needed. This method has already been used for oil engines in several 
instances, but in gas engines it has found next to no application, probably mainly 
because of the added complication and the danger of explosion inherent in the use 
of such separate mixing chambers. That the employment of independent mixing 
chambers is of actual benefit in the process of combustion has been shown in some 
tests on a 4 H.P. Deutz engine made by Petreano with one of his mixing devices 
(see Fig. 396, p. 266). 1 Diagrams taken when the apparatus was connected in all 
cases show a more nearly vertical combustion line and a greater area developed 

than those taken during ordinary operation without previous mixing of charge. 

Combustion, completed at the inner dead center when the superficial area is least, is 
accompanied by a smaller consumption of cooling water, since the heat loss to the 
latter during expansion is less than it was before. To prove this Petreano also 
conducted tests without changing the cooling water in the jackets, that is, cooling 
by vaporization. After working for ten hours with 90% alcohol, the test being 
interrupted three times and distributed over three days, the amount of water that 
had to be replaced was .312 gallon, after operating with gasoline four hours without 
interruption, the quantity amounted to .65 gallon, while after 2| hours operation with 
70% alcohol the temperature of the water had not gone beyond 176° F. The diagrams 

were from beginning to end uniformly good, but nothing is stated regarding 

horse-power, etc. 

Forming the mixture outside of the cylinder in a special chamber is therefore a 
method which should receive some attention in practice, especially if, as seems probable, 
the rotative speeds of gas engines should continue to increase as they have done. 

If it is not practicable to assist diffusion by extending the time allowed, it only 
remains to accelerate it by artificial means, that is, besides using a liberal excess 
of air, to employ some mechanical method of producing the same result. Excess air 
can only be used within narrow limits, the last scheme on the other hand deserves 
the special attention of the designer. The means at present adopted to mechanically 
aid in the diffusion of the gases making up a charge usually depend upon the principle 
of increasing the diffusion areas by minute subdivision of both kinds of gas. This 
method has been considered in greater detail in connection with the mechanical 
details used to carry it out, p. 195. Another scheme serving the same end, that 
of “agitating” the mixture by means of changing direction and cross-section of 
ports, stirring up the charge by sending a vigorous current of air or gas into the 
combustion chamber, etc., deserves to be much more widely tiled than is at present 

*tll0 C&S6. 

In the Diesel constant-pressure engine, the fuel mixture is mechanically agitated 
even during ignition and combustion by forcing a fine stream of air, under a pressure 


i Z. d. V. D. I., 1887, p. 170. 



556 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


of from 90 to 150 lbs. higher than that existing in the cylinder, into the charge 
and promoting the general distribution of this secondary air supply by appropriate 
form of injection nozzle, etc. In the explosion engine, in which the charge should be 
uniformly diffused at the moment of ignition, the mixture is artificially agitated to 
any extent only during the suction stroke, during the compression stroke the agitation 
amounts only to that incident to the operation. An exception to this rule is found 
in the case of certain 2-cycle engines in which the gas is forced into the cylinder 
while the air is being compressed. The fact that, on account of the short period 
available for the formation of the mixture in these engines, their charges do not 
show less, but rather better, uniformity than those of the older 4-cycle engines,, 
points clearly to the beneficial action of agitation upon the constituents of the 
mixture. 

Korting clearly proved this in some tests made, on one of his 4 H.P. vertical 
gas engines, taking indicator cards at widely varying engine speeds and determining, 
from the course of the combustion line the time of explosion or the inflammability 
of the charge. 1 The ordinary indicator diagram does not admit of such determinations, 
because the events discussed occur at the moment of minimum speed of indicator- 
drum. Korting therefore modified the drum by moving a strip of paper under the pencil 
at constant speed (in this case about 24.5 ins. per sec.) and thus obtained pressure-time 
diagrams, the curve showing the pressure changes as a direct function on time. 
Two of Korting’s diagrams, reduced to three-fourths original size, are shown in Figs. 
689 and 690, the former for the highest speed, n=155, and the latter for the lowest,. 



Tenths of ecs. Tenths of Secs. 

Fig. (j8c. 

n = 62. These diagrams show that as far as this engine is concerned (inlet port 
located directly under the cylinder head) the absolute duration of combustion decreases 
as the engine speed increases. Thus for n = 62, Fig. 690, the time of explosion was 
.045 sec., while for n= 155, Fig. 689, this time had decreased to .02 sec. The time 
of combustion therefore varies in nearly the inverse ratio with the r.p.m. or piston 



Tenths of Secs. 


Tenths of Secs. 


Fig. 690. 


speed, in fact it decreases at a little greater rate than the latter increases. Korting 
seeks the reason for this fact in the greater degree of mechanical agitation that the 
mixture receives at higher engine speed. The entering velocity of the charge for 
n = 62 was about 62 ft./sec., while for = 155 it was about 156 ft./sec. This 
increased velocity probably not only increased relative velocities of the gas particles 


1 Z. d. V. D I., 1888, p. 261. 










COMBUSTION IN THE GAS ENGINE 


557 


among themselves in the inlet passage, but in the cylinder as well, and thus promoted 
diffusion and improved the uniformity of the charge. 

It is quite possible, however, that the result shown may be partly due to the 
influence of other circumstances. For instance, higher piston speed is usually accom¬ 
panied by higher exhaust temperature and hotter cylinder walls, and this may serve 
to pre-heat the charge and render it more inflammable. The weight of the burned 
gases remaining in the clearance decreases with the higher temperature, and since 
Korting kept the weight of the fresh charge the same at all speeds, it follows directly 
that the charge grew richer with every increase in the speed of engine. This in 
itself increases the inflammability of the mixture. It should be noted that in these 
tests, the speed variations were not obtained by exceeding the normal speed for 
which the engine was designed, but by starting it a speed below normal, and hence 
the volumetric efficiency of the suction stroke was nearly the same on all tests. 
The case is entirely different when with the increase of speed the normal speed of 
the engine is exceeded. In such cases the suction as well as the exhaust resistance 
losses increase, less fuel charge is drawn in, and more of the burned gases remain 
in the clearance space. This in turn decreases the charge weight and affects the 
purity of the mixture. Under such circumstances, instead of obtaining, as Korting 
did, decrease in the time interval required for combustion, we shall, of course, fin d 
considerable retardation. 

A close examination of the pressure-time diagrams reveals the fact that they 
apparently controvert Korting’s views regarding the “explosion port,” see p. 547, 
while they confirm those of Otto. The full size diagram Fig. 691 (n = 62) shows 

■9at 



near the beginning of the explosion line that the original curve ab has a temporary 
break at the point c; after this the line shows the full rise to d. This indicates that 
the main explosion was preceded by a partial explosion, and the events were probably 
as follows: a to b, explosion in the explosion port, b to c, piercing of the main 
charge by the resulting needle flame. It takes a certain time for the flame to spread 
.and for the effect to become apparent, hence the only slight increase of pressure from 
b to c. After the main charge is ignited at c the pressure rises very rapidly. This 
break in the line occurred, in all of the eight diagrams given by Korting, although 
the indication is not as strong at the higher speed, presumably because the flame 
•spreads more rapidly under the changed conditions. 

Very little reliable data is available regarding the velocity of flame propagation in 
gas-engine fuel mixtures. Many attendant circumstances, such as nature of the fuel, 
purity, composition, temperature and pressure of charge, shape and size of combustion 
chamber, kind and location of igniter, etc., have a marked effect upon combustion 




558 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


phenomena, and render the experimental results that have been obtained on this point 
so uncertain that they are of little value to the designer. In order to obtain values 
at all practical, some experimenters have used apparatus which reproduced moie nearly 
the conditions existing in practice, and especially in place of the usual glass-tube 
models used in laboratories, have employed as combustion chambers metallic vessels 
approaching much more nearly to real working cylinders. In this way at least a 
few reliable comparative figures have been obtained. 

Clerk carried on tests of this nature very soon after the appearance of the new 
Otto engine. 1 He used for this purpose a cylinder, without piston, 7 ins. in diameter 
and 8i ins. long. An igniter was screwed through the bottom cover, while connection 
with an indicator was made through the top cover. The charge fired consisted of 
pure fuel-air mixtures, and the pressure increase was recorded on the drum of the 
indicator which was revolved at constant speed. The curves thus obtained were hence 
similar to Kortmg’s pressure-time curves and allow of the interpretation of the various 
events on the basis of time. The bundle of diagrams shown in Fig. 692 for instance 



Fuel Mixture 
Ratio Gas to Air= 
1 to 14 
1 to 13 
1 to 12 
1 to 11 
1 to 9 
1 to 7 
1 to 6 
1 to 5 
1 to 4 


rftT d Wi ‘ h T tUKS ° f u air and ° ldham coal « as ' At ‘he moment of ignition 
It T atmospheric P ressure > ‘he temperature being from 60 to 66° F. 

mixture in ratio 1 ^14 ^airth” 1 ™* 11 ^ taWe accom P an y i ng Fig. 692, that the 
Mixtures in wUch the ^V IoWest ’ while 1:6 gave the most rapid combustion. 

bCh . th ® ratl ° 1:6 was exceeded, that is, in mixtures still richer the 
inflammability again decreased, as shown by curve i. From diagrams a and * we 
find a mean time of explosion of .05 sec. Since the height of the cylinde/was 81 ins 
the maximum velocity of flame propagation was accordingly about !8™ ft )2’ 

tM “ the rati ° 2:5 sh °™ d velocities nearly three time" 

t figure and developed explosion pressures exceeding 120 lbs per sq in Clerk also 

eqZ U to S 32M“ a F Ua ind na tt m r “i” 1 ^ 011 temperature from diagrams g and h as 

&£££&*: temp “’— g that the heat 

mixture” a The tests with compressed air-illuminating gas 

m xtures The test cylinder used had a diameter of 6 ins., while the length was 

“ - 1SZ £ £ ‘Z7ZL*z 


1 See Clerk, f< The Gas Engine,” ed. 1886 

'Z. dw V. M., 1886, p. 875. 















































559 


COMBUSTION IN THE GAS ENGINE 

obtained with the richest mixture (1:5.42). Increasing the compression pressure from 
atmosphere to 30 lbs. apparently increases the time of explosion from .01 to .0125 
sec. This retardation of the rapidity of explosion grows less marked as the mixture 
grows leaner. Thus for a ratio of 1:7.5, the diagrams of the second row show that 
while the time of explosion under atmospheric pressure (diagram a) is .032 sec., the 
time has increased only to .036 sec. (diagram c), when the mixture is originally 



compressed to 37 lbs. If, as in the case of Clerk’s tests, the velocity of flame 
propagation is computed from the dimensions of the test cylinder and explosion port, 
we obtain the following results for the richest and the leanest mixtures: 


-p, • i . , f Without compression 

Rich mixture 

I With compression . . . 


74.0 ft/sec. 
59.3 


Lean mixture 


| Without compression 
l With compression. . . 


23.0 “ 

20.5 


The retarding influence of compression therefore seems to gradually decrease or 
to entirely disappear, as the mixtures grow leaner. That the kind of fuel used plays 
an important part in this phenomenon has, however, already been pointed out, pp. 536 

and 557. 

In order to study the influence that the form of the combustion chamber may have 
upon the nature of the combustion, Korting also made ignition tests in a tube 2.04 



ins. in diameter and 16.4 ft. long. These tests showed that the combustions in the 
pipe are very irregular, being in some cases explosive and accompanied by violent 
noise, while in other cases they were slow and uncertain. It was also shown that the 
charge could be completely ignited in much shorter time when the ignition started 
from the middle of the length of the pipe than if the charge were ignited from one 











































560 THE GAS ENGINE FUELS AND COMBUSTION IN GAS ENGINES 


end. In all cases, however, the development of maximum pressure took a longer 
time in the long tube than in the short cylinder. Both the shape of the chamber 
and the location of the igniter therefore are of importance in the process of com¬ 
bustion. Theoretically, of course, the spherical form of chamber with the igniter at 
the center is the most desirable, because in this case the distances that the flame 
must travel to complete the ignition are of the same length in all directions 

and at the same time the shortest possible. In actual practice of course this 
ideal form of chamber can never be carried out, the next best shapes - are the 
hemispheres or the short cylinder. In any case the igniter should be placed in the 
axis or center line of the chamber, as near as possible to the core of the com¬ 

pressed charge. 

Another type of diagram which, besides the pressure-time diagram above mentioned, 
may serve to clearly bring out ignition and combustion events is the so-called 
“distorted” diagram Fig. 695, introduced by Prof. E. Meyer. To obtain such a 
diagram, the indicator reducing motion is offset with respect to the engine crank 

so as to bring the explosion line in the middle of the card, when the drum speed 
is the highest. With greater clearness even than the diagrams of Korting, the 

diagrams of Fig. 695 show the partial explosion preceding the main explosion, a 




phenomenon which is hard to understand except by the action of the explosion port 
in the cylinder head. Meyer himself explains its occurrence accordingly. The diagrams 
were taken from a Deutz horizontal kerosene engine. 1 On the basis of the data 
available, Professor Meyer found that the velocity of combustion of the kerosene 
vapor-air mixture was from 32 to 62 ft/sec, which is from 2'to 3 times the value 
found for lean illuminating gas mixtures. Another example of the distorted diagram, 
taken from a Haselwander engine, is given in Fig. 551. 

A further means of studying what happens during ignition and combustion is 
afforded by the “double-stroke” diagram which the writer has used since 1899 
in connection with high speed engines, especially automobile machines. If the indicator 
is operated from the half-time shaft or from any other lay shaft running at one-half 
the speed of the crank-shaft, and if the reducing motion is so set as to bring the 
end of the compression and exhaust strokes in the middle of the travel, we will 
obtain an unusual diagram, the theoretical shape of which is shown in Fig. 696. In 
this diagram the pressure lines of the I and IV, and of the II and III strokes are 


1 Z. d. V. D. I., 1895. p. 987. 


















COMBUSTION IN THE GAS ENGINE 


561 


recorded side by side. The combustion line is taken at the moment of maximum 
drum speed, as in the distorted diagram, and on that account shows any variation 
with great clearness. If the combustion were instantaneous (explosive) at the dead 
center, this line would rise vertically. On account of slow combustion, however, it 
inclines to the right, is as indicated in broken line in Fig. 696. The two double¬ 
stroke diagrams shown in Figs. 697 and 698 were obtained by the writer from a 




3 H.P. gasoline engine running at 1400 r.p.m. The left side of the diagram records 
the compression stroke, the right side the explosion and exhaust strokes. Since, in 
such high speed engines, the spark is made to jump as early as one-half of the 
compression stroke, and since the speed of the indicator drum increases rapidly toward 
the center, the compression and combustion lines merge into eaeh other. The com¬ 
pression curve alone (no ignition) is however clearly shown in Fig. 698, which records 
several miss strokes and late ignitions. 1 Diagrams of this kind can of course not be 
used to compute engine horse-power. 


1 The writer’s pamphlet, “ Konstruction and Betriebsergebnisse von Fahrzeugmotoren,” Berlin, J. 
Springer, contains a number of other diagrams of the kind together with a description of the indicator used 


















































I 



























































































































































































APPENDIX 


A. THEORY 1 


I. Synopsis of Thermodynamics 

1. Heat and Temperature and their Units. Radiant heat is considered as 
produced by transverse vibrations of ether and the latter are considered due to 
rapid vibrations of the molecules of the radiating body. Temperature is one of the 
several manifestations of heat which makes the existing degree of molecular movement 
(degree of heat) sensible externally. Heat and temperature are therefore different 
conceptions. As a measure of temperature we employ another manifestation of heat, 
that of the expansion of bodies. Of the latter, mercury and alcohol are the mosr, 
used in ordinary temperature meters (thermometers). The subdivision of the distance 
between the two fixed points on the thermometric scale, freezing point and boiling 
point, is arbitrary. Besides the now quite extensively used scale of Celsius, which 
divides the distance into 100 equal parts, we further have the scale of Reaumur, 
which uses 80 equal divisions, and the Fahrenheit scale, almost exclusively used in 
this country and in England, in which the distance is divided into 180 equal parts. 
The latter is the oldest of the three. The conversion from one thermometric scale 
to another may be made by the use of the following equations. 

C =- 32°]; R=\c=\[F- 32°]; 

4 9 oo 


F = 32°+^C = 32°+^R. 

The lower limit of commercial utility in the ordinary mercury thermometer is at 
about —40° C.= —40° F., the upper limit is usually not much above 250° C. =482 F. 
When the space above the mercury is filled with nitrogen, brief readmgs^ up to 
350° C =662° F. become possible, but the indication is uncertain above 28U C.-5db t. 
Carbon dioxide thermometers may be used up to 550° C. = 1022° F. For the measure¬ 
ment of still higher temperatures, pyrometers are employed. Their action depends eit er 
upon longitudinal expansion or change of form of solid bodies, upon the expansion 
of air or upon the electro-motive force generated by thermo-elements. Anothei way 


■ A brief and elementary guide for those readers of 
Thermodynamics or Thermochemistry. 


the book not thoroughly familiar with either 


563 







564 


APPENDIX 


of determining very high temperatures is to compute them from the specific heat of 
metallic bodies by calorimetric methods. 

Suppose for instance that a wrought iron sphere of G pounds weight which had 
been exposed some time to the action of the exhaust gases just outside of the exhaust 
valve of a gas engine were quickly dropped into a non-radiating vessel containing 
W lbs. of water, then the temperature of the latter will increase from tf to some value 
t 2 °. If the mean specific heat of wrought iron (or any other metal) was c for the 
temperature range, ti° to t 2 °, the temperature of the exhaust gases must have been 


t=t 2 + - 2 degrees. 

According to Weinhold, we have for wrought iron 


c = .105907 + .00003269 (h + 1 2 ) + .00000001108[*i 2 + 1 2 2 + (h + 1 2 ) 2 ]. 

Note that here t is in Centigrade degrees. 

For temperatures less than —40° C. or F., alcohol thermometers, whose field 
of application ranges from —100° C. to +78° C. (—148° F. to +172.4° F.) are used. 

Just as temperature is expressed in degrees, the quantity of heat is expressed in 
heat units (B.T.U., British thermal units). A heat unit is the quantity of heat 
required to raise the temperature of one pound of water under atmospheric pressure 
through 1° F. [In the metric system the heat unit (kilogram-calorie, or calorie for 
short) is the quantity of heat necessary to raise 1 kilogram of water through 1° C. 
The gram-calorie, sometimes called the small calorie, is equal to t ^ 0 of the kilogram- 
calorie. 

2. Pressure, Density and Specific Weight or Gravity. The “heat carriers” in 

gas engines are gaseous bodies or mixtures of them, simply called “gas” for short. 
Every gas, owing to its tendency to expansion, exerts upon the walls of the vessel 
enclosing it, a pressure, the magnitude of which is expressed either by the height 
of a column of liquid maintained in equilibrium by it or by the pressure exerted 
upon unit area of the enveloping surface. The instruments used in the first system 
of measurement are called manometers, the liquid being mercury or, for very small 
pressures, water; those used in the second system are designated simply pressure 
gauges. The unit of pressure is the “atmosphere,” in English units equal to 14.7 lbs. 
per sq.in., or, expressed in metric units, equal to 1.0333 kg. per sq.cm., at sea level. 
For technical purposes this value of 1.0333 kg. per cm. has throughout Germany been 
replaced by a unit equal to 1 kg. per sq.cm., equivalent to 14.2 lbs. per sq.in., which 
gives the conversion factor used throughout this book. The following table expresses 
the relation existing between the units when expressed in various ways. 


True Atmosphere. 


Metric Atmosphere. 


Kg. per sq.cm. Lbs. per sq.in. 


1.0333 14.70 


Inches, 

Mercury 

(Hg.) 

29.92 


Feet, 

Water. 

33.91 


Inches, 

Kg. per sq.cm. Lbs. per sq.in. Mercury 

(Hg.) 

1.00 14.20 28.74 


Feet, 

Water. 

32.75 


A barometer reading of 29.92” (=760 mm.) is usually taken as the standard 
pressure, while 32° F. ( = 0° C.) is taken as the standard temperature. The two 







THEORY 


565 


together determine the standard condition of a gas (unless for certain purposes another 
standard set is chosen, see p. 6). 

In practice the pressure of a gas is usually stated in pounds above atmospheric 
pressure, that is, it is really what might be called the “over-pressure” (gauge 
pressure). In thermodynamic investigations, on the other hand, the pressure is always 
used as absolute pressure, that is, measured above a vacuum. 

Density (d) of a gas is defined as the weight of a given volume of the gas 
divided by the weight of an equal volume of pure, dry air. The weight of 1 cu.ft. 
of a gas at 32 F. and 29.92” Hg. (14.7 lbs. per sq.in.), is called the specific weight 
(r). Since 1 cu.ft. of pure, dry air under the conditions stated weighs .08072 lb., 
we have that 


r = .08072£, 


( 1 ) 


or 


( 2 ) 


.08072' ' ' 


There are two other ways of expressing the value of y. From p. 569 following, 
we have, eq. (19), that 



in which v — volume of 1 lb. of gas in cubic feet, P = pressure in lbs. per sq.ft., and 
P = absolute temperature in °F. =461 + temperature as measured. R expresses the work 
done by 1 lb. of gas when heated through 1° F., the pressure P remaining constant. 
R is a constant for any one gas. We may write 



1 =J P 
v TR ’ 



Now under standard conditions, P=14.7 1bs. per sq.in. = 2117 lbs. per sq.ft.., and 
T =461 +32=493°, therefore 


_ 2117 4.29 

r ~ 493 R ~~ R 


(la) 


—lbs. per cu.ft., 


The second expression for the value of y is derived directly from the law of 
Avogadro. According to this 



(15) 


in which m is the molecular weight of the gas (see p. 580). 

Under standard conditions (32° and 14.7 lbs.) v 0 cu.ft. of gas will weigh 


Go = yvo =-08072dvo lb, 


(3) 











566 


APPENDIX 


If however, a gas whose specific weight is y, instead of being under normal 
conditions, is under a pressure of p pounds absolute or 6 inches of mercury absolute, 
at a temperature of t°, v cubic feet of the gas under the conditions given will weigh 

C - rv P 493 - rv — 493 lbs . . (4) 

(T ~ r 14.7 m+t r 29.92 461 +t * * * U 

The density of the gas under the same conditions will then be 


S' = d 


493 


493 


14.7 461+< 29.92 461 + <' 


(2a) 


Under ordinary conditions, atmospheric air is considerably lighter than dry air 
on account of the water vapor carried. 

3. Expansion, Absolute Temperature, and Specific Heat. If the temperature of 

a gas is raised from 32° to t°, while the pressure remains constant, the original 
volume v will expand to 

vi=*41+a(t-32)] f .(5) 


or, if the temperature at the beginning is t'°, instead of 32°, 

l+atf-32) x 

* 1 = *l+5((~32j. (5o) 

a is called the coefficient of expansion at constant pressure. For changes at constant 
volume, the pressure changes according to a similar law. The pressure coefficient is 
slightly greater than the volume coefficient a , although the difference is practically 
negligible. For all perfect gases the value of a is nearly the same, that is 


“=493 = - 00203 - 


A volume of gas v x at a temperature t° and pressure b inches of mercury, will 
therefore, when reduced to standard conditions, occupy a volume equal to 

4936 v\b 

V °~ Vl 29.92[493 + {t —32)] = 29.92[1 +.00203 (t -32)]. .^ 

This equation is valid whether heat is supplied or withdrawn. Assuming now 
that it were possible to cool a gas down so far as to make its volume zero, eq. (6) 
shows that the temperature would have to be decreased 493° below 32°, the freezing 
point. This point, —461° below the zero of the Fahrenheit scale, is called the absolute 
zero, and absolute temperatures are measured from it as a starting point. All computa¬ 
tions in thermodynamics are based upon absolute pressure and absolute temperatures. 

« The specific heat (capacity for heat) is the quantity of heat (in B.T.U.) which 
must be supplied to one pound of a given body to raise its temperature 1° F. In the 
case of solid bodies, the specific heat varies with the temperature; in gases it is also 
dependent upon the pressure. A given body of G pounds weight requires a quantity 
of heat equal to 

Q = G ) cdt B.T.U., 


(7) 

















THEORY 


567 


to raise its temperature from t\ to t 2 , through a range of t°, the value of the specific 
heat c not being constant between the temperature limits given. In general, however, 
the variation in the value of c is very small and in many cases yet unknown, hence 
it is usual to assume a mean value of c, in which case eq. (7) reduces to 


Q = Gct B.T.U. 


(7a) 


If G pounds of one fluid, having a temperature t and a specific Heat c, is mixed 
with G' pounds of another liquid, having a temperature t’ and a specific heat c', we 
may write the resulting temperature t m , according to Richmann’s rule, equal to 


Gct+G'c't' . . , 2 (GW) 

tm= s~i , or, in general, Un = 


Gc+G'c' 


S(Gc) 


( 8 ) 


The specific heat of gases at constant volume (c v ) is always considerably smaller 
than that at constant pressure (c p ) and we may write for all perfect gases 


Cp , 2x 

— = x>l; Cp = —-— 

Cv m{x— 1) 


Cv — 


m{x— 1) ’ 


Cp Cv — — AfJ. 
m 


(9) 

( 10 ) 


The conversion of the specific heats c v and c v rom the weight (pound) basis to the 
volume (cubic feet) basis is made by means of multiplying by the specific weight 

r = 35H17 (see p * 581) so that 


c p = yep, and c v = ycv. 


Table 143 

SPECIFIC HEATS AND CONSTANT R FOR PERFECT GASES 


Name of Gas. 


Air, (dry and pure). 



Specific Heat. 

II 

§ lie 

R 

Per Pound. 

Per Cu.ft. at 32° F. 
and 29.92" Hg. 

Cp 

Cv 

Cp 

c v 

h 2 

3.43 

2.43 

] 


r 

770.2 

.0 2 

.217 

.153 

1 


| 

48.30 

,n 2 

.245 

.174 

\ .0191 

.0136 

1.41 \ 

55.08 

CO 

.245 

.174 

1 



55.21 

.NO 

.23 

.163 

J 


l 

52.47 


.238 

.169 




53.34 

h 2 o 

.48 

.37 

.0241 

.0185 

1.300 

85.87 

.CO, 

.20 

.155 

.0245 

.0189 

1.293 

35.13 

c 2 h 2 

[.346] 

[-27] 

[.0251] 

[.0196] 

[1.281] 

59.42 

,ch 4 

.593 

.468 

.0265 

.0208 

1.270 

96.47 

:c 2 h 4 

.40 

.33 

.0354 

.0292 

1.210 

55.19 

.C,H a O 

.453 

.400 

.0576 

.0508 

1.113 

33.56 

•C 0 h 6 

.33 

.305 

.0713 

.0659 

1.082 

19.80 








































568 


APPENDIX 


Table 143 contains the specific heats for all of the so-called perfect gases likely 
to be used in the gas engine industry. Mixtures of gases having the same number of 
atoms have the same volume specific heats as the constituent gases; thus for a 
mixture of di-atomic gases, for instance, we have, from Table 143, 

c p = .0191 and c v = .0136. 

If a mixture consists of G x , G 2 , G 3 , ..., parts by weight or of v x , v 2 , v 3 , . .., 
parts by volume of various gases, so that 2 ((r w ) = l, or 2 (v n ) = l, the respective 
specific heats of the mixture will be 


Cp —2(CrCp)n, Cv —2 (GrCvjn'f .(11) 

Cp = 2 (vCp)n, Cv — 2 (vCv)n. .(11a) 


4. The Relations between Volume, Pressure, and Temperature. The Equations of 
State or Condition. For a certain given weight of a gas, the product of the volume v 
by the pressure p, or the product of the volume v by the density d, is a constant as 
long as the temperature remains constant, that is 



vp=v x pi (Law of Boyle) .... 

.(12) 

and 

vd = v x d x — 1.0. 

.(12a) 

Hence 

v\ p d . 

.(13) 


As stated, eq. (13) only holds good as long as the changes are “ isothermal/' 
i.e., as long as the temperature remains constant. If, on the other hand, the 
temperature during the change increases from T to T x , we will have 


v T p T d T 

— = 7 fr, or - = or 3 - = ■=-, 

V\ 1 \ p 111 Ol 1 1 


(14) 


according to which the 


at constant 


f pressures (densities) 
1 volumes 


are to 


' volumes 

. pressures (densities) 
each other as the absolute temperatures. (Law of Gay-Lussac.) 

The first two equations of (14) can also be derived in another way. Thus from 
eq. (5) we have, if a volume v is heated from 32° to some temperature t at constant 
pressure, that the final volume will be 


*>i = o[l+a(*-32)] = v[l +.00203(4 -32)] = t^l+^^.(15) 

Similarly, if during the same range of temperature the volume remains constant, we 
will have a final pressure equal to 


Pi = P[l +«(£ —32)] = p[l + .00203 (£—32)] = p^l + 


/ —32 


493 


(16) 










THEORY 


569 


Substituting in eqs. (15) and (16), 

T = 493, and T x = [493 + (f -32)]=461+*, 


we have as before 


or 


'H-'El an d ^ = 

v - T ’ and p t’ 

T x , T x 

vi = v^- and p x = p-^. 


(17) 


If the temperature at the beginning of the change had been some temperature t' 
instead of 32°, v x (and by analogy also p x ) would have been directly found from 
eq. (5a). 

Suppose that one pound of a gas, of volume v, is heated through 1° at a constant 
pressure of P pounds per sq.ft. If the final volume is v X) and we designate the work 
done by the pound of gas in expanding from v to v x , against the pressure P by R, 
we shall have 

P(v x -v)=R, 


or 


Now from (17) 


hence 




n_T x _ 

V T’ 

Pv{^-l)=Pv(^)=R. 


But since we are heating only through 1°, T x — T=\, hence finally 

Pv = RT, . 


• (18) 


r is a constant within ordinary limits for any one gas, but it should be noted that 
very low temperatures or very high pressures cause a variation in R, and (18) is 
then no longer strictly applicable. Eq. (18) is a combination of the law of Boyle 
with that of Gay-Lussac and may also be written 


=R= constant, 


or 


v 

v x 


P X T 

PT X 


(19) 


(19a) 


For values of R, see Tables 143 and 147. 

If a gas is compressed in a non-conducting cylinder, its temperature is increased 
from T at the beginning to some temperature T x . This in itself, according to the law 
of Gay-Lussac, causes an increase in the pressure. The latter is hence increased in 
two ways, first, by the pressure on the piston and, second, by the effect of the heat 
of compression. Consequently the pressure exerted by the gas must increase faster 









570 


APPENDIX 


than it would without the temperature rise. The reverse takes place when a volume 
of gas is expanded from v to V\, in which case the work of expansion is done at the 
expense of the heat contained in the gas. This action decreases the temperature which 
decrease is accompanied by contraction, and hence the pressure falls faster than if 
the temperature during the expansion had remained constant. 

According to the law of Poisson, we have for adiabatic changes of state, during 
which no heat is either supplied to or withdrawn from the gas, that 


pv x = piVi x ; . . . ( 20 ) 


P_ = MV 
Pi W ’ 


. . ( 21 ) 


and 


v 

Vi 



. (22) 


in which p and v refer to the original state of the gas, p\ and V\, to the final state, 
and x=C P +C v , the ratio of the specific heats of the gas in question. 

These equations are best solved by means of logarithms, thus 


= anti-log log , .... .(21a) 

and ^- =anti - l0 «(? l0 «^). (22a) 


By combining eqs. (20) to (22) with eq. (14) of Gay-Lussac, we may derive expres¬ 
sions giving the increase or decrease of absolute temperature with adiabatic increase 


or decrease of pressure. We have 



i 


(23) 

and 

V 

v x ~ 

m ■ 

• • (24) 

(25) 

and 


(£)*■ ■ 

• • (26) 


If the original state of a gas is defined by p, v, and T, the end conditions, after an 
adiabatic change, are 


Pi = V = anti- og [log p+z ^log j ; . . 

Pi = p{^) X ^anti-log [log p+^-^(\og yj ; 

vi ~ v (£)‘ = anti M l0g M( l0 M] ; • • 
Vi=v (y x Y 1 = anti_log [ log v + x ~[( log ] ; 

Ti — T (^j = anti-log [log T + (x — 1) log j ; . 

ri=r (l)~ =antUog [ logr+ V( l0g 7)]'• 

The expression (log — ), etc., may also be written (log v — log iq), etc. 


(27) 
(27a) 

(28) 

(28a) 

(29) 

(29a) 









THEORY 


571 


Table 144 

AVERAGE VALUES OF c p , c Vl AND x FOR AIR AND PRODUCTS OF COMBUSTION 



Cp 

Cv 

Cp 

— = x 
Cv 

I 

X 

x — 1 

X 

x —1 

1 

x —1 

x — 1 

X 

Air, pure and dry. 


.2375 

.1684 

1.410 

.709 

.410 

3.493 

2.493 

.291 

For products of combustion 

f 1 to 6 

.268 

.198 

1.356 

.738 

.356 

3.809 

2.809 

.263 

from illuminating gas-air mix- • 

1 to 9 

.259 

.189 

1.370 

.730 

.370 

3.703 

2.703 

.270 

tures in the ratio of 

11 to 12 

.254 

.184 

1.380 

.725 

.380 

3.632 

2.632 

.274 


It is simpler in many cases to determine the end temperature T\, from the general 
relation Pv = RT. It should be remembered in this connection that P represents 
absolute pressure in pounds per sq.ft., and v is the specific volume in cu.ft. of one 
pound of gas. If instead of v the actual volume V of the gas is given in cubic feet, 
so that the weight is G pounds, we will have 


V 1 

* o r 

and from VP=GRT, we then get 

r -?iYl 
Tl GR 


(30) 


5. Examination and Construction of Pressure-Volume Curves. The theoretical 
pressure-volume curves are of importance in thermodynamic investigations because 
they indicate what the pressure and volume changes should be in theory, and conse¬ 
quently form a basis upon which to judge the real changes occurring in practice. 
To construct or draw the curves, the rectangular 
system of coordinates, ox , oy, Fig. 699, is employed, 
the x-axis (abscissas) representing a scale of volumes, 
and the ?/-axis (ordinates) a scale of pressures. 

These scales may be chosen to any convenient 
length per unit. From o, the origin, lay off a dis¬ 
tance v = oa equal to the original volume, and at 
a erect a perpendicular of length ab, so that ab = p 
is the original pressure. The point b then repre¬ 
sents the original state, designated by p, v, T, of a 
certain quantity, say one pound of the gas. If now 
we allow the gas to expand to say 5 times its original 
volume, the point c, at a distance equal to 5v = 5ao 
from the origin along the x-axis, will represent 
the final volume Vi. The length of the ordinate 
through c, however, depends upon the equation of condition or state according to 
which the gas expanded. If, for instance, by means of a proper supply of heat, 
the pressure had been maintained constant, we will have cd==o6, so that Pi=p, and 
the curve would have been a straight line bd parallel to ac. This is known as a 
constant pressure line or an isobar. The heat supplied has of course raised the 
temperature from T to T h and the final state of the gas at the point d is given by 





p, v lf TV 





























572 


APPENDIX 


If during the expansion just so much heat is supplied as is required to maintain 
the temperature constant at T , the pressure p would have decreased, according to 

V 

Boyle’s law, eqs. (12) and (13), and at every stage of the expansion Pi = P —• I f at a 

number of points between a and c, ordinates are erected equivalent to the correspond¬ 
ing pressures, as determined from the above equation, and if the end points of these 
ordinates are connected by a smooth curve, the curve will be the isothermal be. The 
final condition at the point e is given by v 1} p h T. 

Finally, if we assume that no heat whatever is supplied to the gas during the 
expansion from v to V\, and that the enclosing walls are impermeable to heat so that 
the change proceeds without either supplying or withdrawing heat (adiabatic change), 
we may apply the laws of Poisson, as stated in eqs. (20) to (22). If the ordinates 
corresponding to any intermediate volumes are computed from these equations, it will 
be found that the pressures decrease much faster than they did for the isothermal line 
and that consequently the adiabatic line bf diverges more and more rapidly from the 
isothermal be. The end condition at the point / is defined by p\, v ly TV 

If we keep in mind that the isothermal line be follows the law while the 

adiabatic line is determined by the equation = (~^j > ^ be seen once that ^he 

greater the exponent x, the more rapidly will the adiabatic approach the rr-axis. If 
the value of the exponent is some other value than x, say 1, the simplest way of 
constructing the curve is by the method of Brauer. 1 From the origin, Fig. 700, lay 
off the angle a below the z-axis, and compute the angle /? for the y-axis according to 
the equation 


1 +tan /?= (1 -1-tan a) n . 


(31) 


From z, the point fixing the original condition of the gas and hence the starting 
point of the curve, draw the vertical zs and the horizontal zw. Through the inter¬ 
sections s and w draw the oblique lines at 45° 
as shown, and from the points si and W\ so 
determined draw the next set of verticals and 
horizontals. Their intersection in Zi determines 
the next point on the curve. Continuing this 
process determines the curve. The net-work 
of construction lines not only determines the 
polytropic line through the point z, but it is 
possible by their use to draw a series of such 
polytropics. A second curve is indicated in 
Fig. 700. The smaller the angle a is taken, 
the closer together will be the points of inter¬ 
section on the curve, and the more accurate 
therefore will be the curve. Table 145 gives several practical values of a and /? for 
exponents from n = l.l to 1.41. 


V' 

, ^ 

\z 



\ 


,7. 

7V f •?£■ -* 

774, V 



\ 



A 7 
















\ 0 



S. 

s„ 



X 


1 

9 ' 



/C*t 




Fig. 700. 


1 Z. d. V. D. I., 1885, p. 433. 




















THEORY 


573 


Table 145 


Exponent n = 

1.10 

1.15 

1.20 

1.25 

1.30 

1.35 

1.41 

Angle a 0 

11.20 

11.20 

11.20 

14.05 

14.05 

14.05 

18.25 

Angle ft 0 

12.35 

13.10 

13.50 

17.55 

18.40 

19.25 

26 .30 

Tan a. 

0.2 = 1:5 

0.2 

0.2 

0.25 = 1:4 

0.25 

0.25 

0.33 =1:3 

Tan /?. 

0.222 

0.234 

0.245 

0.322 

0.337 

0.352 

0.497-1:2 


For n=1.0 (isothermal line) the pressure-volume curve becomes a common 

hyperbola. 

Instead of constructing the pressure-volume curve from given data we can of 
course go through the reverse process and from given pressure lines, as for instance 
those given by the engine indicator, determine the condition of the gas with respect 
to v, p , and T. Since the exponent in the equation for the adiabatic change is 


equal to x — — } we know at once that, if an actual pressure-volume expansion line for any 

Cv 


given gas shows an exponent n different from x for this gas, the change was not 
adiabatic. If in such a case n>x, heat has been withdrawn during the change, while, 
if n <x, heat must have been supplied. On the other hand, the opposite relations 
hold for the compression line. 

In practice, however, an expansion or compression line may appear to have an 
exponent n = x and still not be a true adiabatic. This occurs when the amount of 
heat withdrawn from the gas during the change is just balanced by the amount 
supplied. Consequently if the expansion line of a gas engine indicator diagram shows 
an exponent equal to x, this by no means proves that the expansion has been 
adiabatic, but rather that a considerable amount of after-burning has taken place. 
The highly heated gases must during the expansion transfer a considerable quantity 


of heat to the cooling water, etc., which quantity can only be restored by after¬ 
burning if the expansion line appears adiabatic, that is, shows n = x. A curve of this 
kind has by some writers been called a false adiabatic. 

From this standpoint we may summarize what is said above as follows: In all 
expansion lines for which n<x, that is, those that are situated above the theoretical 
adiabatic curve, the supply of heat is greater than the simultaneous withdrawal of 
heat, while the reverse is the case for all expansion lines for which n>x. 

The deductions concerning the process of expansion made on the lines of Fig. 
699, also apply without change to the compression of all perfect gases. The adiabatic 
compression line therefore rises more rapidly than the isothermal compression line 
starting from the same point, and for all lines lying f 
above the isothermal the greater the exponent n in 
the equation pv n = const., the steeper the slope of the 
curve. 

The exponent n for that part of a polytropic curve 
lying between the end points 1 and 2, Fig. 701, may 
be found from the equation 


n = 


log pi -log p 2 
log T >2 — log V\ ' 


(32) 



in which p and v are as indicated in the figure. In this case n is considered constant 
for the entire curve. In fixing the points 1 and 2 care should be taken to see that 





















574 


APPENDIX 



they are located on the expansion line proper and not on the end of the combustion 
or the beginning of the exhaust line respectively. If it is desired to determine the 
variation of n along the curve, divide the latter into a sufficient number of parts 
(5 to 10) and from the values of p and v for the end points of each one of the 
partial lengths, say between the ordinates x and y, n may be computed from 


log p x -log Vv 

X/V log Vy —log Vx 


(32 a) 


An accurate investigation of gas engine diagrams always calls for a subdivision 
of the expansion or compression lines, as above indicated, into as many parts as 
practicable, sometimes on an enlarged scale, because the exponent n, especially in the 
case of the expansion line, often changes very considerably. (See p. 38, Part I.) 

6. Relations between Heat and Work, and the Mechanical Equivalent of 
Heat. A gas, expanding when heated, does external work, and conversely, doing 
work upon a gas, as compressing it, generates heat. There is a certain definite 
relation between the unit of work (foot-pound) and the unit of heat (British Thermal 
Unit) and it has been found experimentally that to generate mechanically one thermal 
unit requires the expenditure of mechanical energy equivalent to 778 ft.-lbs. Con¬ 
versely, one B.T.U., when completely utilized, can do work equivalent to 778 ft.-lbs. 


Hence, 1 B.T.U. = 778 ft.-lbs. is the mechanical equivalent of heat, while 1 ft.-lb. = ~— 


B.T.U. may be called the heat equivalent of work. The latter value is usually 
designated by A, so that 


A=-^ = .Q01284 B.T.U. 


The first accurate determination of the mechanical equivalent was made by Joule, 
who found at that time ~~772 ft.-lbs. This figure is generally known even to-day 
as Joule’s equivalent. 1 

Heat supplied to a body of gas has three effects: it increases the mobility of 
the molecules, which manifests itself on the one hand in greater amplitude of vibration 
(temperature) and on the other tends to change the relative position of or the distance 
between the atoms (internal energy), the two effects together transforming a certain 
amount of the heat into internal work. The third effect of supplying heat is to 
increase the volume of the gas or its pressure, and this represents external work. 
If the two parts of the internal work U are designated by W and J respectively, 
and the external work by L, we may write for an infinitesimal supply of heat that 

dQ = A(dW +dJ+dL)=A(dU+dL) .(33) 

Eq. (33) expresses mathematically the first law of Thermodynamics. If for dJ+dL 
we write dH, so that 

dQ = A(dW+dH). 


1 Regarding the uncertainty of this value, which varies from 772 to 782, see p. 2, Part 1. 






THEORY 


575 


the first law may also be stated as follows: “The heat supplied to a body is partly 
used to raise the temperature and partly to increase the volume or the pressure 
of the same.” 

Assume now that we have one pound of any one of the perfect gases, with a 
volume at the start equal to v cu.ft., an absolute pressure of P pounds per sq.ft., 
and an absolute temperature of T degrees, enclosed by a piston in a non-conducting 
cylinder. If we next supply this body of gas with a certain quantity of heat equal 
to Q B.T.U., the effect will depend upon the action of the piston and may be as 
follows: 

(а) Volume Constant. If the piston is held in its original position, (v = constant), 
the heat Q supplied will raise the temperature of the gas from T to T x , and the 
pressure from P to Pi, according to the equation 

T x = T+^- = T^ .(34) 

Cv i 

T 

Conversely, having given Ti and P X = P 7 ^, we can find Q from [see also eqs. (27) 

to (29)], 

Q = c v (T 1 -T)=^(P 1 -P) B.T.U.(35) 

Since v remains constant, the heat supplied did no external work. 

(б) Pressure Constant. If the gas during a movement of the piston is furnished 
with a quantity of heat Q, sufficient to keep the pressure P constant, the temperature 
in the cylinder at the end of the expansion will be 

T^T+Q-^T 1 ^ .(36) 

Cp V 

The external work done by the piston in its outward movement is equal to the 
pressure P into the change of volume, that is, 

L = P(v x —v) ft.-lbs., or L = R(Ti—T) ft.-lbs.,.(37) 

and the heat used will be 

Q=c v {Ti-T) = ~-(vi—v) B.T.U.(38) 

(c) Temperature Constant. If the supply of heat is so controlled that the temper¬ 
ature does not change during expansion, the final pressure will be 

P x = P~ lbs. per sq.ft. 

The external work done will be 

L = 2.3026 RT log ^ =2.3026 RT log ^ ft.-lbs.,.(39) 

n v 

L = 2.3026 Pv log - = 2.3026 P x v x log ^ ft.-lbs. 

° V V 


or 


. . (39a) 






576 


APPENDIX 


Finally, the heat required is 

Q =='2.3026 ART log ^ = 2.3026 ART log ^ B.T.U.(40) 

The factor 2.3026 represents the ratio of'natural to common logarithms, and disappears 
if the former are used in place of the latter. 

(d) No Heat Supplied during Expansion. If the gas expands without gain or 
loss of heat (adiabatically) from the original state (p, v, T) to the volume v h the 
heat received is of course equal to zero. The change in pressure and volume is 
determined by eqs. (21) and (22). The work done is 



The expressions developed above under (c) and ( d ), when used in the proper 

sense, also apply to the compression of a gas. As before, v, P, and T refer to the 

original and v\, P\, and T\ to the final conditions. In this case, however, Q is the 
heat generated by the compression, that is the heat to be withdrawn, and it should 
be noted that it refers in all cases to one pound of gas. 

7. Cycles. A cycle is defined as a series of successive pressure and volume 

changes in a body of gas, during which heat is supplied to and withdrawn from the 

latter in such order and quantity that the final con¬ 
dition of-the gas is the same as the original. The 

pressure-volume lines of a cycle therefore form a 
closed polygon, but the individual enclosing lines 
may follow any law. Fig. 702 may serve as an 
example. Suppose that from the point 1 one pound 
of air (gas), having a volume v x cu.ft., absolute 
pressure Pi lbs. per sq.ft, and an absolute tem¬ 
perature T, expands isothermally to the point 2. 
In order to maintain the temperature at T, we 

must supply a certain quantity of heat Q' B.T.U. 

From 1 to 2, the pressure has fallen from Pi to 

P 2 lbs. per sq.ft. The external work done during 
this change was L' ft.-lbs., from eq. (39), repre¬ 
sented in the diagram by the area F' = 1 2III. The heat supplied along 1-2 

therefore is 

Q' = AL' = AF' B.T.U.(42) 





Next assume that the expansion from 2 to 3 is adiabatic, so that the external work 
must be done at the expense of internal energy in the gas. The temperature then 
falls from T to T n the pressure from P 2 to P3 lbs. per sq.ft., while, acc ording to 
eq. (41), the external work done is L, ft.-lbs., equivalent to the area F / = 2 3IIIII. 
The quantity of heat withdrawn from the gas to do the external work is 

U l = AL t = AF l B.T.U. . . 


. . (43) 












THEORY 


577 


The quantity of heat Q supplied between 1 and 2 has during the expansion of 
the gas consequently done a total amount of work equal to (L'+L,) ft.-lbs., repre¬ 
sented by the area 12 3 III I. 

The piston next starts on its return stroke and first compresses the gas isother- 
mally to the point 4. During this change a certain quantity of heat Q" must be 
withdrawn from the gas in order to keep the temperature constant at T 3 in spite of 
the fact that P3 increases to P 4 . The work of compression L" may be computed from 
eq. (39). It is represented by the area F" = 3 III IV 4 and its heat equivalent is 

Q" = AL" = AF" B.T.U.(44) 

If from the point 4 the compression proceeds adiabatically, without supply or 
withdrawal of heat, the gas will return to the original pressure Pi and the original 
temperature T when the volume is again reduced to v\, provided that the point 4 or 
the volume has been properly chosen. The external work L u done upon the gas 
between 4 and 1 is determined from eq. (41). It is represented by the area 
F„ = 4 IV I 1 , and the internal heat (heat of compression) generated is expressed by 

V„-AL„ = AF„ B.T.U.(45) 


The hatched area 1 2 34 = P represents the amount of external work done by the gas 
in passing through this cycle and is equal to 

L= '(L'+L, —L" —L„) ft.-lbs.(46) 


Eq. (46) expresses the difference between the quantity of heat Q' supplied and the 
quantity Q" withdrawn, hence 


Q’-Q" _Q 
A A 


ft.-lbs. 


(47) 


In eq. (46) since the external work done by the gas along 2 3 is equal to the 
work done upon the gas along 4 1, 

Lj = L /r 

Hence the total work done also equals 

L = L'-L" .(47a) 


From eq. (39), however, 

L’~ 2.3026 RT log and L"= 2.3026 RT, log 


therefore 


L = 2.3026 RT log ^ -2.3026 RT, log 











578 


APPENDIX 


But for the adiabatic changes — =—, hence 


Vi v 4 


L = 2.3026 R (T - T t ) log ^ = 2.3026 log ^.(48) 


(48) 


In this equation the factor 2.3026 RT log — corresponds according to eq. (40) to 


Q' 

the quantity of heat supplied and substituting this we finally have 



(48a) 


This equation expressed in words states: that the theoretical external work L 
done in a reversible cycle depends not only upon the quantity of heat Q' furnished 

T 

to the cycle but also upon the ratio of the temperature limits. 

The cycle just described is generally designated “reversible” because the individual 
changes could also be carried out in the reverse order with the same final result. In 
the special case above outlined, in which the limiting curves consist of two isothermals 
and two adiabatics, the cycle is the ideal {Carnot cycle) because for the given 
temperature limits it shows the theoretically best utilization of heat, that is the 
thermal efficiency i) t is the maximum. If, on the other hand, the heat has been 
supplied or withdrawn at varying temperature the cycle is called poly tropic (according 
to Lorenz). Finally, we also distinguish “ closed,” and “open” cycles, depending upon 
whether the same body of working fluid always goes through the various changes of 
state or whether the working fluid is subject to constant renewals. 

The thermal efficiency of a cycle is the ratio between the amount of heat trans¬ 
formed into external work divided by the total heat supplied to (or used by) the 
cycle. According to whether the heat supplied or the external work done refers to 
the theoretical cycle or to the actual indicator diagram, we distinguish 


Theoretical thermal 
Indicated thermal 



The ratio of these two corresponds to the efficiency of the combustion process and 
may also be considered to represent the “card factor,” that is 



If as before Q f is the heat supplied, Q" the heat discharged, so that Q'-Q"=Q 
is the heat transformed into external work, we have in general 


Q'-Q" Q _AL 
Q' Q' Q' ’ 


and for the ideal cycle in particular, as will easily be seen from eq. (48a), 


(49) 









THEORY 


579 


In the case of the Carnot cycle, therefore, the thermal efficiency depends only 
upon the range between the temperature limits, that is upon the fall of temperature. 

If the indicated work of one cycle is L { , while the heat supplied is Q, the 
indicated thermal efficiency will be 


Vi= 


U 

L 


AU 

Q 


(50) 


For the method of determining L { , see p. 8, Part I. 

8. “ Heat Weight ” and Entropy. From the derivation of eqs. (49) and (49a) 
it follows that for the ideal cycle we may write 


&__T_ 
Q"~T / 


hence also 


q_ 

T 


Q" 

t; 


(51) 


This ratio of the quantity of heat supplied or discharged to its absolute tem¬ 


perature is usually called the “heat weight,” following Zeuner’s example. Eq. (51) 
may therefore be expressed in words as follows: “ In an ideal cycle the heat weights 
supplied and discharged are equal to each other.” This is the second laic of thermo¬ 
dynamics. 

Writing eq. (48a) in the form 


Q=AL = ^(T-T,), 


(52) 


Q' 

and keeping in mind that is the heat weight and T — T t the temperature drop in 


the cycle, we may also state the second law as follows: “In an ideal cycle that part 
of the heat ( Q ) transformed into external work is equal to the product of heat 
weight by the temperature drop.” 

Since Q' represents the heat supplied and Q" that discharged, so that the 
quantities are of opposite sign, we may write 


14 = 0 . 


If now we assume that any given closed cycle is divided into an infinite number 
of elementary Carnot cycles, we must therefore have also for the summation of all 
these elementary cycles that 


s |=0 or 



Clausius has given the expression ff the name “Entropy.” With this conception 

the second law of thermodynamics states that “ In any given closed cycle the 
entropy (sum of the heat weights supplied and discharged) is equal to zero.” 

By plotting entropy with temperature we obtain the so-called entropy diagram 
which is an invaluable aid in the study of the operation of internal-combustion 
engines (see p. 10 and 40, Part I). 





580 


APPENDIX 


II. Fundamental Principles of Thermochemistry 

1. Atoms and Molecules and their Weights. A given body may be divided both 
chemically and mechanically. The infinitesimally small particles of an elementary 
material that can be divided no further either mechanically or chemically are known 
as atoms. It is assumed that in any given material these atoms are closely associated 
in certain groups, and such smallest possible atom-groups into which the material can 
be subdivided without affecting its chemical constitution, are called molecules. 1 
Expressing this explanation in other words we may say that a molecule is the smallest 
weight necessary for the existence of a substance, while the atom is the smallest 
conceivable weight with which the chemical elements form combinations. 

The atoms of the various elementary materials have a weight, called atomic 
weight (e), which, although impossible of direct determination, is definite and fixed. 
This weight is usually referred to that of hydrogen (the lightest element) as unity, 
and therefore expresses how many times the atom of a given body is heavier than 
the atom of hydrogen. The atomic weight of oxygen is e= 15.95. If this figure is 

16 

taken at 16.0, the atomic weight of hydrogen becomes = 1.0032. For practical 

reasons chemists have of late years found it better to make computations with oxygen 
e=16.0 than with hydrogen e=1.0. 

The sum of the weights of all of the atoms of a molecule is known as the 
molecular weight ( m ). Since one molecule (H 2 ) of hydrogen consists of two atoms, 
its molecular weight is m = 2. If based upon oxygen, however, (0 2 )=32, the molecular 

32 

weight of hydrogen is w = = 2.0064. Chemical computations are usually made with 

such units as the gram-molecule or the gram-molecule-volume, the numerical values 
for which are given below. 

The number of molecules contained in a given volume of a body constitutes its 
mass, 2 while the relation between mass and volume is the density, ( d ) of the material. 
According to the law of Avogadro, equal volumes of all gases at the same pressure 
(density) and temperature contain the same number of molecules. From this we derive 
the following important conclusions: 

(а) The molecule-volume of all gaseous bodies is the same. 

(б) The density of all gaseous bodies is in direct proportion to their molecular 
weight m. 

In general the density of gases is referred to that of air as unity, that is d 
expresses how many times heavier a given gas is than an equal volume of air under 
the same conditions. 

In conformity with the law of Avogadro, above stated, it has been found that the 
pound-molecule-volume is equal to 358.17 cu.ft. Mathematically expressed this means 


1 To get a clear conception of these statements, a gaseous body may be considered as a dust-cloud, in 
which the individual dust grains are made up of the finest elementary particles of some elementary 
substance. The dust grains then represent the molecules and their constituent particles the atoms. 

O 

2 Mass (M) and weight (G) of a body are proportional to each other, but not equal, since Af = ——. 

oZ.Jd 

The mass of a body is constant, but the weight depends upon the acceleration due to gravity, that is 
upon the geographic location. 




THEORY 


581 


that, if m represents the molecular weight as before, the volume of one pound of any 
of the perfect gases under standard conditions, the so-called specific volume v, is equal 
to 


358.17 ,. 

V — — - cu.ft., 

m 


( 1 ) 


and conversely that the weight of one cubic foot of any of the perfect gases under 
standard conditions is equal to 

1 m 


r=~=; 


: lbs. 


V 358.17 

From this equation the last column of the following table has been computed. 


( 2 ) 


Table 146 

CHEMICAL CONSTANTS FOR PERFECT GASES 


Kind of Gas. 


Hydrogen.H 2 

Oxygen.0 2 

Nitrogen.N 2 

Carbon monoxide.CO 

Nitric oxide.NO 

Carbon dioxide.C0 2 

Air, dry. 

Water vapor.H 2 O 

Acetylene.C 2 H 2 

Methane.CH 4 

Ethylene.0 2 H 4 

Alcohol.C 2 H 6 0 

Propylene.C 3 H 6 

Benzol.C e H a 


Atom 

Number. 


2 

2 

2 

2 

2 

3 

3 

4 

5 

6 
9 
9 

12 


Atomic 

Weight. 


1 

16 

14 


For carbon (C 2 ), e = 11.965~12; m = 23.93~24; o' = .820; and r= 0668. 


Molecular Weight. 

Density 
Referred 
to Air= 1.0 

d 

Weight 
per Cubic 
Foot at 
32° and 
14.7 lbs. 

r 

Approxi¬ 

mate, 

H 2 = 2. 

m 

Accurate for 

H 2= 2 - 

m 

0 2 = 32. 

m 

2 

2.0 

2.006 

.0693 

.005588 

32 

31.9 

32.0 

1.1056 

.0892 

28 

28.00 

28.00 

.971 

.0783 

28 

27.93 

28.00 

.967 

.0781 

30 


: 30.40 

1.031 

.0838 

44 

43.87 

44.00 

1.519 

.12262 

29 

28.88 


1.0 

.08069 

18 


18.0 

.623 

.0503 

26 

25.94 

26.01 

.915 

.0724 

16 

15.97 

16.02 

.554 

.0446 

28 

27.94 

28.02 

.974 

.0781 

46 


! 46.03 

1.601 

.13072 

42 

41.91 


1.451 

.1171 

78 

77.82 

78.04 

2.695 

.2177 

0; and y ■■ 

= .0668. 





2. Elements and their Combinations and Symbols. Substances that can not 
be subdivided into simpler ones are called elements. Two or more elements may umte 
to form one or more chemical compounds either by the direct combination of their 
molecules or by a rearrangement resulting in what would correspond to the substi¬ 
tution of the atoms of one of the elements in the molecule of the other. 

Chemical combinations are entered into with ease, with difficulty, or not at a , 
depending upon the chemical relationship (affinity) of the elements concerned. If the 
chemical combination takes place it always occurs without loss of matter (weight) but 
is usually accompanied by a decrease in the number of molecules and hence contraction 

° f SP Elemente U Tnd chemical combinations are designated by the first letter of their 
(usually Latin) names, the number of atoms per molecule being at the same time 








































582 


APPENDIX 


defined by numerical subscripts. The chemical symbols of all of the gases likely to 
be used in the gas engine industry are given in Table 146. The symbol for hydrogen, 
for instance, is H 2 , which states that the molecule of hydrogen consists of two atoms. 
Leaving out the atom number (in this case 2) gives to the symbol a double meaning, 
the letter H then not only signifies hydrogen in general but also a definite chemical 
quantity, that is, one atom of hydrogen, which in chemical relations has only one 
definite value. The same applies of course to the chemical symbols of all of the 
elements. 

One atom of carbon (C) combines with one atom of oxygen (0) to form a 
diatomic molecule of carbon monoxide (CO). In this case the combination formula 
(CO) lacks the atom number because the number of atoms in the union is definitely 
stated by the finking of the two symbols. The symbol for carbon dioxide is C0 2 , 
which states that it is a combination of one atom of carbon with two of oxygen, and 
one molecule of C0 2 therefore consists of three atoms. The same number of atoms 
is contained in water vapor (H 2 0), ammonia (NH 3 ) has four atoms, methane (CH 4 ) 
five, benzol (C 6 H 6 ) twelve, etc. If a group of similar combinations is to be indicated 
by one general symbol, the atom number is replaced by the subscript n. For instance, 
C„H 2n designates a combination of n atoms of C with 2 n atoms of H. In the case 
of ethylene (C 2 H 4 ), therefore, n =2, while for hexylene (C 6 H 12 ), w = 6, both belonging 
to the group of hydrocarbons designated collectively by C w H 2re . The symbol C n H 2n+2 
states that a molecule contains n atoms of C to 2w+2 atoms of hydrogen, as "for 
instance methane, CH 4 . 

The determinations of volumes or weights of chemical combinations is done by 
means of their molecular weights, the same as for the elements. If the molecular 
weight m is not given it may be easily computed from the atomic weights e of the 
elements in the combination. 

Examples. Carbon monoxide has a molecular weight equal to 27.93, consequently 1 cu ft 
from eq. (2) will weigh 

27.93 

r = 358 J 7 = -° /81 P ° Unds * 

The same result would have been obtained by starting with the weight of a standard cubic 
foot of hydrogen, and remembering that the densities of the perfect gases are to each other as 
their molecular weights, as follows: 


T— .005588 X—-— = .0781 pounds. 


0^3 molecule of carbon dioxide is composed of one atom of C (whose atomic weight is 

- =11.965) and two atoms of 0 (whose atomic weight is — 2 ~ = 15.95);. its molecular weight 

therefore is (sum of the weights of the three atoms) equal to 11.965 + 2X15.95 = 43.865 One 
cubic foot of C0 2 consequently weighs 

43.865 

r- 358a7 - ' 12262 P° unds ' 

and its density is (air=1.0), 


.12262 

.08069 


= 1.519. 





THEORY 


583 


One molecule of ethylene consists of 

2 atoms of C, of weight = 2X11.965 = 23.93, 
and 4 “ “ H, “ “ =4X1.0 =4.00, 

hence the molecular weight of C 2 H 4 = 27.93 


27 93 

and 1 cu.ft. of C 2 H 4 weighs ~ = .0781 lbs. at 32° F. and 14.7 lbs. abs. pressure. 

358.1 / 

Any hydrocarbon combination of general formula C„H ri has a molecular weight of 

M 0 =11.96n + ln=(11.96 + l)n~13n .(3) 

in which 11.96 = atomic weight of C and 1.0 = that of H. For the so-called heavy 
hydrocarbons of general formula C n H 2/l we have similarly 

M 0 = 11.96n + 2n=13.96n~14n,.. (4) 

and the density of the latter group referred to air as unity will be 


14n 

,08069 


(5) 


If any chemical combination is composed of several elements Ei, E 2 , E s , .. . , 
whose atomic weight is e x , e 2 , e 3 , , and the number of atoms in which is 

«i, n 2 , w 3 , . . . , respectively, the molecular weight of the combination is 

m = n\e\ + n 2 e 2 -\-n 3 e 3 -\- . . . = 2(ne).(6) 


3. Combining Weights and Volumes, Atomic Heat, Diffusion and Dissociation. 

According to the law of conservation of matter, the elements and their weights remain 
unchanged during any chemical change. A combination of different gases therefore 
has exactly the same weight as the sum of the weights of the constituent gases. 
Chemical reactions in general only take place in definite weight proportion and m 
the case of the gases usually also in simple volume proportions. he smallest weight 
(or volume) with which any given element will still combine with another element 

is known as its combining weight (or volume). . 

In gas engine practice it is often most important to know the volume ratios in 
which given gases combine. This is most easily determined by writing down the 
chemicaf combination formulas of the gases concerned, the number of molecule* of 
each gas in the combination will then give the volume, as the following examples show. 


(o) Combustion of C to CO: 

1 molecule of C 2 + l molecule of 0 2 = 2 molecules of CO, 


hence 

1 volume of C + l volume of 0 = 2 volumes of CO 








584 


APPENDIX 


(b) Combination of hydrogen and chlorine gas: 

1 molecule of H 2 + l molecule of Cl = 2 molecules of HC1, 

hence 

1 volume of H + l volume of Cl = 2 volumes of HC1 (hydrochloric acid). 

(c) Combustion of H to water vapor: 

2 molecules of H 2 + l molecule of 0 2 = 2 molecules of H 2 0, 

therefore 

2 volumes of H + l volume of 0 = 2 volumes of H 2 0. 

(d) Combination of C and H to form methane, CH 4 : 

1 molecule of C 2 + 4 molecules of H 2 = 2 molecules of CH 4 . 

therefore, 

1 volume of C+4 volumes of H = 2 volumes of CH 4 . 


The term atomic heat means the product of the specific heat of an element by its 
atomic weight. The atomic heat is a nearly constant quantity, having about the same 
numerical value (6.0 to 6.5) for all of the elements. It follows from this that equal 
numbers of atoms require the same quantity of heat to raise the temperature the 
same amount. Diffusion is not a chemical combination but is simply the process by 
which the molecules of liquids or gaseous bodies form perfectly uniform mixtures, 
there being no chemical reaction of the elements (or molecules) upon each other. 

In atmospheric air for instance we find diffused nitrogen and oxygen. Air is therefore 
not a chemical compound. Diffusion of gases is unlimited, as long as the gases do 
not react chemically, but it takes place in different gases with varying rapidity. 
The measure of the power of diffusion is called the diffusion coefficient , which represents 
the “ diffusion length,” that is, the distance traveled by one gas in another in one 
second. The experimental data available on this point is rather scarce and somewhat 
unreliable as far as practice, as distinguished from laboratory work, is concerned. 
Thus for carbon dioxide and air at 32° and 29.92" Hg., various tests show the 

diffusion coefficient to vary from .053 to .059 in. per second. In general, the lower 
tlie density of the gases and the greater the mobility of the molecules, the more 

rapid the diffusion. For that reason an increase of temperature and mechanical 

agitation will aid the process, while compressing the gas retards it on account of the 
increase in density. According to the investigations of Loschmidt, the diffusion 
coefficient varies inversely as the density (pressure) of the gases and nearly directly 
as the square of the absolute temperatures. 1 In compressing a gas, however, although 
the increase of presssure decreases the diffusion coefficient, the simultaneous decrease 
of volume decreases the diffusion distance in the same ratio and hence the latter effect 
balances the former. The result therefore is that pressure or density has no influence 
upon the diffusing power of gases. 


Meyer, Kinetische Theorie der Gase, p. 250. 



THEORY 


585 


It has been shown that the relative proportion of the gases in the mixture has 
no sensible influence upon the diffusion process. 

The field of every chemical action is limited by an upper and a lower tempera¬ 
ture. Either above the upper limit or below the lower limit, chemical reaction ceases, 
and the combinations formed decompose into their constituents, at least above the 
upper limit. This decomposition process, which thus depends upon the temperature, 
is known as dissociation, and the temperature at which it makes its appearance is 
called the dissociation limit. At temperatures exceeding 3600° F., for instance, water 
vapor (H 2 0) breaks up into hydrogen and oxygen, and carbon dioxide into carbon 
monoxide and oxygen. Since the development of heat in an internal-combustion engine 
is the result of chemical reaction, it follows that the combustion temperatures of the 
mixtures must not be allowed to exceed the upper dissociation limit, otherwise the 
development of heat would be interrupted at the upper limit until the temperature has 
again fallen below it. An increase of pressure usually raises the upper dissociation 
limit, a drop of pressure consequently lowers it. 

Every combination of oxygen with other elements is known by the general term 
oxidation, and is generally accompanied by a decomposition, development of heat, or 
other phenomena. The most important oxidation process is combustion. The reverse 
process, that is, the extraction of oxygen from oxygen compounds, is known as 
reduction. While oxidation produces heat, reduction absorbs it. Among the practical 
reducing agents are carbon, hydrogen, certain hydrocarbon compounds, etc. Carbon 
(C) for instance, is oxidized to C0 2 (carbon dioxide). If now the C0 2 formed is 
led through an incandescent layer of carbon, it is reduced according to the equation 
C+C0 2 = 2C0 (carbon monoxide). These alternate actions are of the greatest import¬ 
ance in the production of producer gas (see p. 593). 

4. Combustion. Most chemical reactions are accompanied by a development of 
heat, although this may not be easily noticeable in some cases. During the combina¬ 
tion of oxygen with certain related elements, however, notably hydrogen and carbon, 
the accompanying chemical energy is so completely transformed into heat that the 
materials concerned either glow or ignite, thus making the process visible to the eye 
and perceptible to the touch. This phenomenon is known as combustion Oxygen 
combinations of this kind only form in definite atomic or weight ratios, which may be 
expressed by chemical equations as follows: 

H 2 + 0 = H 2 0, that is, two atoms of hydrogen combine with one atom of oxygen to form 
one molecule of water vapor containing three atoms. 

C + 0 2 = C0 2 , that is, one atom of carbon combines with two of oxygen to form one mole¬ 
cule of carbon dioxide containing three atoms. 

C + 0 = C0, that is, one atom of carbon combines with one of oxygen to form one mole¬ 
cule Of carbon monoxide containing two atoms. 

After substituting the atomic weight for the atomic number (see table 146, p. 581) in 
above relation, these may be written. 

2 lbs. of H + 16 lbs. of 0 = 18 lbs. of H 2 0 1 M 0 l eC ular weight of H 2 0 is therefore m— 18 

1 lb. of H + 8 lbs. of 0= 9 lbs. of H 2 0 J 

12 lbs. of C +32 lbs. of 0 = 44 lbs. of C0 2 \ Molecular we ight of C0 2 is therefore m = 44 

1 lb. of C+f| lbs. of 0 = 3.666 lbs.ofC0 2 ' 

12 lbs. of C + 16 lbs. of 0 = 28 lbs. of CO \ Molecu i ar weight of CO is therefore m=28 

1 lb . of C + if lbs. of 0 = 2.333 lbs. of CO J 


586 


APPENDIX 


The combustible elements are not usually available in the pure form, but are 
found in combinations or mixed with other and indifferent elements. 

In order to start combustion there is required a certain minimum temperature 
known as the ignition temperature. The latter may be produced chemically (by means 
of glowing or flaming materials, etc.) or mechanically (by means of pressure, friction, 
etc.). The temperature of ignition for any given element apparently varies with 
attendant circumstances, and is therefore not definitely fixed. For hydrogen it is 
about 1020° F. and for methane and carbon monoxide in the neighborhood of 1200 6 
F. The ignition temperature for carbon seems to be about midway between these 
values. The low ignition temperature of hydrogen explains why gas engine fuel 
mixtures high in hydrogen can stand only a moderate compression. The combustion 
temperature is of course much higher than the ignition temperature, although the former, 
on account of the interruption of the process of oxidation caused by dissociation, 
practically never reaches the theoretically possible value. Hydrogen gas, for instance, 
if burned in pure oxygen without dissociation, would show a combustion temperature 
of from 11 700 to 12 600° F., while carbon under the same conditions would even 
reach 18 000° F. In an actual case, on the other hand, neither combustion, on account 
of dissociation, shows much over 3600° F. In general, a non-luminous flame burns 
at considerably higher temperature than one that is luminous. The reason for this is 
found in the heat consumed by the solid, carbonaceous particles which give the light 
in the flame, and their incomplete combustion. A non-luminous flame is produced 
when the fuel gas before its ignition is furnished with so much air (oxygen) that its 
combustion is instantaneous and complete (Bunsen burner). 

The quantity of heat which one pound of the combustible produces during the 
union with oxygen is called the heating value and differs considerably for the various 
fuels. Thus one pound of hydrogen, when completely burned, will develop 62 000 
B.T.U., one pound of carbon 14 544 B.T.U., and one pound of sulphur 3966 B.T.U. 

A fuel is burned completely when in the oxidation process all carbon is changed 
to carbon dioxide, all hydrogen to water vapor, and all sulphur to sulphur dioxide. 
If the combustion does not attain this “ maximum degree of oxidation” the heat 
developed may be considerably less. Thus carbon, when burned to carbon monoxide 
instead of carbon dioxide, develops only 4446 B.T.U. per lb. Chemically this inter¬ 
relation between heat of combustion and degree of oxidation is indicated as follows: 


H 2 + 0 = H 2 0 (liquid)+68.90 cal. per gr.-mol. or 61520 B.T.U. per lb. 


c + o 2 = co 2 

+96.96 

: 6 ( 

14544 

c+o =co 

+ 29.56 

( i ( 

4446 


Any combustion is proved complete when the products of combustion, both solid 
and gaseous, show no combustible or unburned constituents. Since in gas engines only 
gaseous fuels (hydrogen, carbon monoxide, and hydrocarbon compounds) are burned, 
not in oxygen, but in air, the exhaust gas must consist of carbon dioxide, water 
vapor, nitrogen, and excess oxygen. 

Theoretically one atom of C requires two atoms of O for its complete combustion, 
while one atom of H only requires one-half atom of O. The theoretically least amount 
of oxygen required to completely burn one pound of each combustible may then be 
found from the atomic weights as follows: 


THEORY 


587 


For Carbon, C 

G =^^=1 = 2.667 lbs., 

1 Z o 


For Hydrogen, H 

lbs. 


Since one pound of air contains .235 pound of oxygen, the least amount of air 
required for the complete combustion of C 2 and H 2 will be 


Gi= 2 ~^ = 11.35 lbs. for C. 
.235 


and 34.04 lbs. for H. 

.235 

If the fuel does not consist of a single element as above, but is composed of 
combinations of the fuel gases mentioned, or combinations of these with other gases, each 
of the latter must be considered in computing the necessary oxygen or air required 
according to its atomic weight e, or its atomic number n. Thus one pound of a gas 
having the composition (C ni H n „O n3 N n4 ) would require a weight of oxygen equal to 

2nie 3 + .5n 2 e 3 —n 3 e 3 32 n x +8n 2 - 16re 3 /j\ 

in which m is the molecular weight of the gas. The burned gases resulting from 
the complete combustion of this gas will consist of 

fc ^ ni(ci+ _ 2 c3) = 44ni ^ Qf CQ .( 8 ) 

m m 

and « 2 (e 8 + -5e > ) _9n i ^ of H 0 .(9) 

m m 

together with the nitrogen and the excess of oxygen in the air used for combustion, 
both of which remain uncombined. By means of these equations the data of Table 
147 for the main constituents of illuminating gas has been computed. 1 The average 
illuminating gas quoted in the last row of the table has the following composition by 
weight: .54CH 4 +. 10C 2 H 4 + .08C 4 H 8 + .05H 2 + .15CO + .08N 2 . 

By use of the columns for s, k and w, in this table, the theoretical amount 
of oxygen required by an illuminating gas, each pound of which contains say 
Gi lbs. CH 4 , G 2 lbs. 0 2 H 4 , G 3 lbs. C 4 H 8 , . . ., may be directly found by substituting in 


the formula 

Oxygen = 2(Crs) pounds.(10) 

Similarly, carbon dioxide and water vapor contained in the burned gas may be 
determined from 

Carbon dioxide = 2 (Gk) pounds.. (11) 

and Water vapor =2 (Gw) pounds.. « . . . (12) 


1 Mainly following Zeuner, Thermodynamik, 2d ed., I., p. 398. 
















588 


APPENDIX 


Table 147 

COMBUSTION DATA FOR THE PRINCIPAL GASES 


Kind of Gas. 

Lower Heating 
Value, H u 

Theoretically Required. 

Products of 
Combustion per 
lb. of gas Contain 

Specific Heat of 
the Gas 

Gas 

Constant 

R 

Oxygen 

(s) 

per lb., 

lbs. 

Air (Z) 

per lb., 

B.T.U. 

per cu.ft., 

B.T.U. 

per lb. 

lbs. 

percu.ft., 

cu.ft. 

Carbon 

Dioxide 

(A) 

lbs. 

W ater 
Vapor 

(u>) 

lbs. 

c p 

c v 

Hydrogen, H 2 

51696 

289 

8.0 

34.0 

2.38 


9.0 

3.409 

2.412 

770.2 

Nitrogen, N 2 








.244 

.173 

55.08 

Carbon monoxide, CO 

4334 

338 

.57 

2.42 

2.38 

1.57 


.245 

.174 

55.21 

Methane, CH 4 

21385 

952 

4.00 

17.00 

9.52 

2.75 

2.25 

.593 

.468 

96.47 

Ethylene, C 2 H 4 

20025 

1584 

3.43 

14 .55 

14.29 

3.14 

1.29 

.404 

.333 

55.19 

Butylene, C 4 H 8 

19508 

3055 

3.43 

14.55 

28.25 

3.14 

1.29 

.404 

.333 

27.51 

Average illuminating gas * 

18200 

700 

3.26 

13.85 

6.66 

2.29 

1.90 

.619 

.473 

110.87 


* This average quality of illuminating gas is much better than that found in the United States, where 
the average B.T.U. per cubic foot is about 575. 


According to the so-called combination formulae, one pound of fuel, consisting of 
C pounds of carbon, H pounds of hydrogen, O pounds of oxygen, and S pounds of 
sulphur, theoretically requires for its complete combustion 


|C+8H+S-0 

.23 


lb.. 


or 


|C+8H+S-0 
.0187 


cu.ft. of air. 


• • (13) 


If the resulting products of combustion consist of k% carbon dioxide, s% oxygen, 
and n% nitrogen, by volume, the ratio of the air actually used to that theoretically 
required will be 

21 

Air excess coefficient = —...(14) 

21 -79- 
n 


If instead of the weight per cent, the volume per cent composition of a hydrocarbon 
compound C ni H n , is given, one cubic foot of the gas will requiie 

&= cu.ft. of oxygen.(15) 

Example. The derivation and use of (15) is best explained by a problem. Suppose that 
the given hydrocarbon gas is C 2 H 4 . This makesn 1 = 2 and n 2 = 4. The gas may be considered 
made up as follows: 


C 2 + 2H 2 = C 2 H 4 , 

1 volume of C + 2 volumes of H = 1 volume of C 2 H 4 , 


hence the combustion of 1 cu.ft. of C 2 H 4 may be considered equivalent to the combustion of 
1 cu.ft. of C (gas) and of 2 cu.ft. of H. 

























THEORY 


589 


(a) Combustion of 1 cu.ft. of C: 

C 2 + 20 2 = 2C0 2 , 

1 volume of C + 2 volumes of 0 = 2 volumes of C0 2 , 

hence 1 cu.ft. of C requires 2 cu.ft. of O. 

(2) Combustion of 2 cu.ft. of H: 

2H 2 + 0 2 = 2H 2 0, 

2 volumes of H +1 volume of O = 2 volumes of H 2 0, 

hence 2 cu.ft. of H require 1 cu.ft of O. 

The combustion of 1 cu.ft. of C 2 H 4 then requires 

©=2 + 1 = 3= ^n 2 +^ cu.ft. of oxygen, 

which proves eq. (15). 


From eq. (15) air required for the combustion will then be 

2 = ^3=4.695^+^ cu.ft.(16) 

The combustion formula for this gas will read 

CniH n2 + ^fti+-^^0 = niC02+^H20, 

or put into words: One volume of CrnH« 2 will bum with volumes of oxygen to 

form n\ volumes of CO 2 and -y volumes of H 2 O, water vapor. 

The method of determining the numerical factors ni and for carbon dioxide and 

water vapor in the above statement will be clear from the example above given. 

Finally, as a general case, suppose another gas has the following composition by 
volume, the letter x with the proper subscript representing the percent by volume 
of the particular gas. Let the composition be 

aqCO + x 2 H 2 + X 3 CH 4 + X 4 C 2 H 4 + x 5 C 2 H 2 + x 6 0 2 + x 7 N 2 + x 8 C0 2 + x 9 H 2 0 = 1 cu.ft. 

Then the amount of oxygen theoretically required by one cubic foot of this gas will be, 

©= p!±^-2 + 2 x 3 + 3x 4 + 2.5X5-x 6 j cu.ft.(17) 

Aa before, the air theoretically required per cubic foot of gas will be 


2=4.695© = 4.695j ^ 1 + 2x 3 +3x 4 +2.5x 5 -x 6 J cu.ft, 


(17a) 







590 


APPENDIX 


If now £' cubic feet of air are actually supplied per cubic foot of gas burned 


(£'>£), the products of combustion resulting will be 

CO 2 = \x\ +£3 + 2 x 4 + 2 .T 5 cu.ft.;.(18) 

H 2 O — [x 2 + 2^3 + 2 ^ 4+£5 +*rg] cu.ft.;.(19) 

N = [.79S'+x 7 ] cu.ft; 0 = [. 212 '- 6 ] cu.ft.( 20 ) 


These equations represent the cubic feet of burned gases actually produced per 
cubic foot of fuel gas. The results may be easily reduced to per cent composition by 
volume if desired. 

The total volume of burned gases, reduced to the original pressure and temperature 
is usually smaller, never greater, than the original volume of the gases before com¬ 
bustion. The same holds true concerning the number of molecules before and after 
combustion. Contraction therefore occurs. For a gas of the general formula C*H n 
the amount of contraction, according to Mollier, may be expressed by 

JV=(^l— parts by volume, .( 21 ) 

while for the second and more complex composition above given it is approximately 
equal to 


, T7 CO + H 2 +C 2 H 2 A u 

JV = - - -* parts by volume.(21a) 

It is here assumed that the water in the products of combustion really escapes in the 
shape of vapor, which is of course always the case in gas engines. 

In this connection it should be noted that the heating value of a fuel containing 
hydrogen depends upon the final condition of the water vapor (water of combustion) 
concerned in the combustion. If the burned gases leave at a temperature exceeding 
212 ° the water vapor carries with it, and renders unavailable as far as any attempt 
at utilization is concerned, the latent heat of vaporization. The heat of combustion 
developed is thus reduced to the so-called lower heating value (HJ. If on the other 
hand the products of combustion before escaping are brought to a temperature below 
212 °, the water vapor is condensed during the time that the heat is being abstracted 
(used) and the latent heat of. vaporization thus returned increases the available heat 
of combustion to the higher heating value (H 0 ). Since one pound of water vapor 
under average conditions, say when cooled to 65° F., gives up about 1100 B.T.U., 
the difference between these two heating values amounts to 


Ho — H m = llOOw,.(22) 

in which w is the weight of water resulting from the combustion of 1 lb. or 1 cu.ft. 
of fuel. 

As far as gas engine practice is concerned, the lower heating value is always the 
one used, because the latent heat of the water vapor does not become available for 









THEORY 


591 


transformation into work except at lower temperatures, and its value for this purpose 
is consequently small. 1 

To compute the heating value of solid and liquid fuels the following combination 
formula (DuLong’s formula), with certain restrictions applicable also to gases, may be 
used. 

H u = [l4500 C + 52200 (u + 4500S -1100H 2 oj B.T.U. per pound. . (23) 

In this equation, C, H, 0, and S represent the part by weight of the elements 
mentioned, while H 2 0 is the water contained originally in 1 lb. of the fuel. 

If 1 lb. or 1 cu.ft. of a gas mixture consists of t lf t 2 , t 3 parts by weight or 
by volume of the individual gases and the heating values of these gases are 
H 1; H 2 , H 3 . . ., respectively, the heating value of 1 lb. or 1 cu.ft. of the mixture will 
be expressed by 

H m = 2(iH) B.T.U. . ..(24) 

If a fuel gas having a heating value of H B.T.U. is mixed with l parts by 
weight or by volume of air, the heating value of the resulting mixture will be 

= B.T.U. per lb. or per cu.ft.(25) 


The fuels in common use in gas-engine practice consist very largely of the hydro¬ 
carbon groups C n H 2 n +2 or C n H 2n . The heating value of such combinations is very 
simply determined from the heating values of their constituents C and H. The 
compound C„H n consists of n atoms of C and n atoms of H. Since the atomic 
weight of C is 12 and of H is 1, we may say that one pound-molecule of C n H n 
weighs 12n+w= 13n lbs. Now lib. of C will develop 14500 B.T.U., and lib. of 
H, 52500 B.T.U. Hence the heating value of the combination will be 


jj (12 X14 500)n + (1X 52 200)n (174000 + 5 2 200)n _ FTU ner lb 

“ ' ~(12 + l)n 13 n ' 1 


(26) 


It will be noted that the atom number finally cancels out, so that every hydro¬ 
carbon compound of the form C n H n will develop 17 400 B.T.U. per lb. Since in this 
case the molecular weight of the combination is 13n, the weight of a cu.ft. of C n H n 

will be 

13n ,, 

r ~ 358.17 lbs '’ 


and the heating value of C«H n is consequently also 

H u = 17400 X= 632n B.T.U. per cu.ft. 

<500.1 / 

Example. For benzol (C 6 H 6 ), n = 6, hence the heating value is 17 400 B.T.U. per lb., or 
632X6 = 3792 B.TU. per cu.ft. 


1 This particular point was for a long time a much discussed question. See the thorough treatment 
given by E. Meyer iu the Z. d. V. D. I., 1899, p. 326, etc., also same, 1903, p. 632. 









592 


APPENDIX 


In the same way the heating values of the heavy hydrocarbons C re H 2re may be 
computed. In this case there are two atoms of H for every atom of C in the 
combination. Hence the pound-molecule weighs 12n + 2n = 14n lbs. Proceeding as above 
for C„H n , we finally get for C n H 2n 


H w = 19 880 B.T.U. per lb.,.(27) 

and H u = 19880X^4v = 778n B.T.U. per cu.ft.(27a) 

O Do. 1 I 


(Note. It should be distinctly stated that the heating values for the hydrocarbons 
computed as above do not in many cases agree with calorimetric determinations, as 
may be seen from the table following. The reason probably is that the above formulas 
do not, and can not, take into account the molecular energy interchanges that must 
take place in the combustion of one of these complex compounds. It is in all cases 

advisable, no matter what the fuel may be, to make calorimetric determinations of 

heating value.) 

In his “ Kalorimetrische Untersuchungen,” Slaby gives the following formula for 
the computation of the heating value of heavy hydrocarbons. This formula in English 
units reads 

Hu = [112 + 18880f] B.T.U. per cu.ft.,.(28) 

in which the atom number n is replaced by y, the weight of a standard cubic foot 

of the hydrocarbon under consideration. Since 


r= 


m 


358.17 


- lbs., 


where m is the molecular weight of the gas, we may also write 

H lt = [112 +52.8m] B.T.U. per cu.ft.(28a) 

This formula not only applies to the individual hydrocarbons, but, with an error 
which does not exceed from .5 to 1%, also to mixtures of the same (as illuminating 
gas), when the latter contain at least 4% of hydrocarbons, have a heating value of at 
least 550 B.T.U. per cu.ft., and when the value of y is quite accurately known. Table 148 
shows how close Slaby’s formula approximates to true results. 

5. Distillation and Gasification. The conversion of the solid into gaseous fuels 
is done either in retorts or in special furnaces, called generators or producers. 1 In the 
first case the fuel is subject to distillation only, those of the hydrocarbons which can 
be vaporized being driven off together with the water at temperatures ranging from 
900 to 1100° F., the fixed carbon remaining behind. This method (dry distillation) 
is the basis of the manufacture of illuminating gas, see Part IV, p. 514. In the 
discussion following, this gas need be no further considered since the consumer obtains 
it as a finished product from the manufacturer. On the other hand the construction 
and use of suitable power gas generators belongs strictly to the province of the gas- 
engine builder. 


1 Concerning the construction of gas-making apparatus, see p. 269, Part II. 









THEORY 


593 


Table 148 


HEATING VALUE OF VARIOUS HYDROCARBONS 


Name of Gas 

Weight in 
Pounds per 
Cubic Foot 
at 32° F. 
and 14.7 lbs. 

r 

Density, 
Air-1.0. 

d 

Heating 
Value per 
Pound, 
Including 
Water of 
of Com¬ 
bustion, 
B.T.U. 

H 0 

Heating 
Value of the 
Water of 
Combustion 
from 1 Pound 
of Gas. 
B.T.U. 

H w 

Heating Value per 
Pound, | Cubic Foot, 
Exclusive 
of 

Water of Combustion 

Heating 
Value from 
Eq. (28a), 
B.T.U., 

H u 

B.T.U. 

H u 

B.T.U. 

H u 

Hydrogen. 

.H, 

.00559 

.0692 

61520 

9824 

51696 

289 


Methane. 

•CH 4 

.04464 

.5530 

23842 

2457 

21385 

952 

957 

Acetylene. 

• C 2 H 2 

.07251 

.8982 

21429 

756 

20673 

1499 

1487 

Ethylene. 

•C++ 

.07809 

.9674 

21429 

1404 

20025 

1584 

1590 

Ethane. 

•C++ 

.08329 

1.0367 

22399 

1966 

20433 

1682 

1696 

Allylene. 

•C 3 H 4 

.11157 

1.3819 

20992 

983 

20009 

2238 

2244 

Propylene. 

c 3 h 6 

.11699 

1.4512 

21224 

1404 

19820 

2318 

2330 

Propane. 

•C 3 h 3 

.12256 

1.5204 

21825 

1786 

20039 

2424 

2435 

Butylene. 

■C 4 h 8 

.15599 

1.9349 

20912 

1404 

19508 

3055 

3069 


The manufacture of power gas in generators also commences with a distillation 
of the solid fuels, but it soon changes from this directly into a gasification process 
which consumes the fixed carbon of the fuel, and all of the fuel is converted into gas 
with the exception of some incombustible ingredients (ash, clinker, etc.). In the 
combustion processes so far discussed, the combustion was assumed complete, that is, 
the resulting gas contained no combustible constituents. On the other hand, the gas 
made in gas producers must be a combustible gas, which first makes it necessary by 
some means or other (proper height of fuel column, right temperature, regulating the 
quantity of air introduced) to render the combustion process incomplete, that is, for 
instance, to burn the carbon to carbon monoxide instead of to carbon dioxide. If 
only carbon and air are concerned in the gasification process, the resulting air gas 1 
can theoretically contain only carbon monoxide and the nitiogen of the air. According 
to p. 586, 1 lb. of carbon burned to CO 2 develops 14544 B.T.U., but only 4446 B.T.U. 
when burned to CO. The carbon monoxide gas produced therefore contains theoret¬ 
ically 14544-4446=10098 B.T.U. per lb. of carbon gasified. This amounts to 69.5% 
of the heating value of the combustion of the solid C to C0 2 . The remaining 30.5% 
are to be found in the sensible heat of the gas as it leaves the producer, and is lost 
if the gas must be cooled before utilization. In this case the maximum possible 
theoretical efficiency of the process is therefore 69.5%. It will be shown below that 
other methods of gasification show higher efficiencies than this, for which reason air 
gas is used in engines only when it can be had at a very little cost as a by-product 
from other manufacturing processes (blast-furnace gas, for instance). 

The chemical equations of the air gas process, according to p. 585, may be 
written as follows: 

C + 0 2 = C0 2 + 96.96 cal. per gr. mol. = 14544 B.T.U. per lb. of C 
C + 0~ = CO +29.56 “ “ = 4446 “ “ C 

CO + 0 = C0 2 + 67.40 “ “ = 10098 “ “ C 


* Other names, such as “generator gas,” have been used for this gas, but the term “air gas is most 
suitable because it draws the distinction sharply between the carbon monoxide gas on the one hand and 
producer gas proper and water gas on the other, which are also made in generators or producers. 




























594 


APPENDIX 


The combustion of 1 lb. of C to CO, according to p. 585, requires 1.33 lbs. of 0, 
according to which 1 lb. of CO gas will develop during combustion 10098 = (1 +1.33) =4334 
B.T.U. The specific weight y of the gas is, from Table 146, .0781 lb., hence 1 cu.ft. 
of the theoretical air gas has a heating value of .0781X4334 = 338 B.T.U. 

Now, in reality, the air gas process is not carried through with oxygen but with 

1 33 

air, the gasification of 1 lb. of C requiring in theory = 5.67 lbs. of air. The total 

weight of gas produced per pound of C will then be 6.67 lbs. The amount of nitrogen 
present is 5.67 - 1.33 = 4.34 lbs., which is diffused through the 2.33 lbs. of CO formed 
r 4.34 

during the combustion. Hence for every pound of CO there will be =1.86 lbs. of 

N, and lib. of air gas will then consist of ^^^p .35 lb. = 35% of combustible 

carbon monoxide and i n L8 ^ =.65 lb. = 65% of non-combustible nitrogen. Since the 
l.U + l.oO 

densities of N and CO are very nearly alike, the same percentages express also the 
volume composition. The heating value of 1 lb. of this gas, since it contains 35% of 
CO, will be, . 35 X 4334 = 1516 B.T.U. One cubic foot of air gas will weigh .35X.0781 + 

6 67 

.65X.0783 = .07823 lb.; 1 lb. of C consequently produces 973 93 = 85 - 3 CU ' ft ’ °* gas ' 

In the production of “water gas” the sensible heat of the gas, which is lost in 
the production of air gas, is largely saved by converting it into latent heat, that is, 
utilizing it in the production of a second combustible gas. For this purpose the coal 
or coke column in a water gas generator is alternately blown into incandescence by 
means of air and then cooled down by passing steam through it. During the “ blow ” 
the carbon in the charge is by means of the oxygen in the air burned very largely 
to CO, and the temperature is constantly increasing; during the “ make,” water vapor 
replaces air and the former, in passing through the incandescent carbon, splits up into 
hydrogen and oxygen. The oxygen set free combines with C to form CO 2 , but the 
latter, when the temperature is sufficiently high (above 1260° F.) is almost immediately 
reduced to CO. The reducing process consumes a large part of the generator heat 
available at the start of the make, and the temperature constantly falls. This heat 
is then returned by a new blowing period, bringing the temperature again up to the 
maximum. The water gas process is therefore carried on with constant interruptions, 
the blowing periods usually taking much more time than the period of gas making 
proper. 

The higher the generator temperature at which the process is carried on, the 
greater the content of CO, and the smaller that of C0 2 in the resulting gas. Below 
from 1000 to 1250° F. the decomposition of H 2 0 with C takes place mainly according 
to the formula 


C + 2H 2 0=C0 2 + 2 H 2 , 


but as the temperature increases, CO commences to outweigh C0 2 , and above 1800° F. 
the only reaction occurring is probably 


C + H 2 0=C0 + H 2 . 





THEORY 


595 


According to the last equation, a given volume of the theoretically perfect 
(synthetic) water gas therefore consists half of carbon monoxide and half of hvdrogen. 
The weight ratio may be determined from the molecular weights as follows: 

18 

1 lb. of C +^2 lbs - of H 2 0 = 2.333 lbs. of CO + .166 lb. of H 2; 
or 1 lb. of water gas contains 2 333 ^16 6 = *° 665 lb ‘ of h y dr °g en > 


and 


2.333 

2.333+ .166 


= .9335 lb. of carbon monoxide. 


The most unfavorable condition that can exist at the end of a gas-making period 
is defined by the following equations: 

1 lb. of C+^ lb. of H 2 0 = 3.666 lbs. of C0 2 + .333 lbs. of H 2 , 


or 1 lb. of water gas contains 3 666 _^ 333 = .0825 lb. of hydrogen, 

and 3 ~Q 0^^^333 = -91?5 lb. of carbon dioxide. 

The amount of heat rendered latent during the decomposition process may be found 

from 

C + H 2 0=CO + H 2 —38.78 cal. per gr. mol. = 5800 B.T.U. per lb. of C, 

in which it is for the present assumed that the water enters the generator already in 
the form of steam. The theoretically perfect water gas has the following heating 

value: 


Due to CO. 2.333 X 4334 =10110 B.T.U. 

DuetoH 2 .166x61 520=10212 “ 

Total. 20322 “ 


Heat supplied during blowing period (rendered latent) 5800 “ 

Difference = heat in 1 lb. of C=. 14522 “ 

The latter item should have been exactly the heating value of C (14 544 B.T.U.), 

since the process is theoretically without loss. The discrepancy, however, amounts to 

less than of 1 %, and is probably due to various approximations in the computation. 

In practical operation, the efficiency obtained in the water gas generator is only 
from 50 to 60%, the remainder is lost in the sensible heat carried away by the gases 
during the blow, through incomplete combustion, radiation, etc. Efficient generators 
of moderate size produce about 32 cu.ft. of gas per lb. of C (coke) gasified, and from 
about 24 to 29 cu.ft. per lb. of good coal. 

The generation of 1 cu.ft. of water gas consequently requires from .042 to .035 lb. 

of coal. The amount of water used is theoretically about .025 lb. per cu.ft. of gas; 













596 


APPENDIX 


in practice, however, the amount used is considerably greater. The average composition 
of the commercial water gas is in volume per cent as follows: 

50% H, 40% CO, 5% C0 2 and 5%N; H„~280 B.T.U. per cu.ft. 

Weight per cu.ft., r = .044; theoretical air required for combustion is 2.45 cu.ft. 
per cu.ft. 

The Delwick-Fleischer method of making water gas differs from the common 
method in that during the blowing-up period the gas mad is largely carbon dioxide, 
not carbon monoxide. As a consequence the carbon is burned completely and a 
greater quantity of heat is stored up for the decomposition of w^ater. This permits 
of a longer gas making period, and the process is less often interrupted. 

The method of making “power or producer gas” 1 is a combination of the air 
gas and water gas processes. As the name indicates, the gas is mostly used for the 
production of power in gas engines. It is made by sending a certain quantity of 
water vapor along with the air through the incandescent column of coal or coke in 
the generator, in which case the oxidizing and reducing actions above discussed, 
between carbon and oxygen on the one hand and between steam and carbon on the 
other, take place practically at the same time. The final product of this process 
must theoretically again consist of carbon monoxide, hydrogen, and nitrogen. The 
introduction into the generator of air and steam combined possesses two main 
advantages. In the first place it serves to bind very largely (render latent) the 
sensible heat of the combustion products, thus raising the efficiency of the generator, 
and, second, permits of uninterrupted operation of the producer. The air, being 
admitted along with the steam, supports the combustion to such a degree, that is, 
keeps the temperature of the producer at such a point that the heat consumed in 
the decomposition of the water vapor is immediately replaced. 

The individual actions and reactions developed above for air gas and water gas 
of course apply directly to the manufacture of producer gas. In order to avoid 
repetition, however, Table 149, published by H. Gerdes, 2 is given. This table shows 
in concise form all of the fundamental chemical relations concerned in the process of 
making producer gas, and by this arrangement gives a clear insight into what takes 
place. 3 

In actual practice the composition of the power gas is somewhat different from 
the values developed in the table, mainly because the theoretical ratio between C 
and H 2 0 is not maintained, a part of the C entering into other combinations. In 
the best practice there are obtained about 75 cu.ft. of power gas having a heating 
value of 146 B.T.U. per cu.ft. from 1 lb. of coal, the heating value of which is about 
14 000 B.T.U. This corresponds to a generator efficiency of about 78 to 80%. By 
means of recovering or more efficiently utilizing the sensible heat of the gases, which 
at present very largely accounts for the 20% lost, this generator efficiency may be 
raised by from 5 to 10%. 


1 This gas also bears the name “Dowson” gas, but with little justice, since the English engineer, Dowson, 
merely developed the first successful scrubbers for the gas, the method of production for which was well 
known before that time. 

2 Glaser’s Annalen, 1902, No. 590. 

3 A more detailed exposition of the theory of the production of power gas is given by Prof. E. Meyer 
in the Z. d. V. D. I., 1895, p. 1524. This investigation of the process is worthy of note for its theoretical 
treatment. 



THEORY 


597 


Table 149 


FUNDAMENTAL CHEMICAL RELATIONS CONCERNED IN THE MAKING OF PRODUCER GAS 



Weight Relations. 

Total 

Lbs. 

Volume Relations, Cubic 
Feet of Gas Concerned in 
the Reaction. 


H 2 

CO 

CO 2 

N 2 

Total 


C 4- H 2 O = CO + Hj 








1. Water gas. 

12 + 18 = 28 + 2 = 

30 

358 

358 



716 

Hence 30 lbs. of C produce 


+ 51840-10: 610 







2576 cu.ft. of producer 









gas consisting of: 


-51840 B.T.U. 
















13.9% of H 2 


C+O+ N = CO + N 







27.8% of CO 

2. Air gas. 

12 + 16 + 16 X 3.24 = 28 + 51.8 = 

79.8 


358 


661 

1019 

6.9% of CO 2 


+ 51840 B.T.U. 







51.4% of N 2 

1 and 2 together. 

24 lbs. of C+18 lbs. of H 2 O +67.8 lbs. of air= 

109.8 

358 

716 


661 

1736 

1 lb. of C therefore pro¬ 


= 840 cu.ft of air 







duces 85.9 cu.ft. of pro¬ 









ducer gas of an average 

3. To replace heat 

IC+ O + N = £C02+ N 







heating value 135 B.T.U. 

losses (assumed) 

6 + 16 + 16 X 3.24= 22 + 51.8 = 

72.8 



179 

661 

840 

per cu.ft. 

Sum. 

30 lbs. of C +18 lbs. of H 2 O + 135.6 lbs. of air = 

182.6 

358 

716 

179 

1322 

2576 



= 1680.3 cu.ft. of air 









The fuel used in practice is not pure carbon, but ithe latter is contaminated with 
earthy admixtures, sulphur, etc. In making computations on any gasification process, 
this fact must therefore be allowed for either by using the correct C-percentage in 
the fuel at the outset, or by multiplying the results obtained for pure C by a factor 
which expresses the percentage of pure C in the fuel used. The impurities in the 
fuel also have a certain effect upon the composition of the gas, increasing the 
proportion of incombustible gases. Among the latter will always be found several per 
cent of carbon dioxide (see p. 520), produced mainly owing to temperature variation 
in the producer (see p. 600). 

The average composition, volume-per-cent, of commercial power gas is about 23% 
CO, 18% H, 1% CH 4 , 6% C0 2 , 52% N. The lower heating value of the gas is about 
Hn~140 B.T.U. per cu.ft., the weight per cu.ft. ^=.0789 lb., while the air theoretically 
required for combustion is 1.1 cu.ft. per cu.ft. 

The ratio of the heating values of the individual gases in a cubic foot of the average 
producer gas above given to the total heating value of the cu.ft. is as follows, from 
Table 147. 

For CO, heating value = .23X338 = 77.8 B.T.U. =56.5% of H w 

For H, heating value = .18X289 =52.0 B.T.U. =37.0% of H u 

For CH 4 , heating value = .01 X952= 9.5 B.T.U. = 6.5% of H M . 

It is evident that the hydrogen content has an important influence upon the 
heating value of producer gas. In the case of gas made from hard coal especially, 
the hydrogen content often accounts for more than one-half of the total heating 
value (see Table 124, p. 520). It is consequently of advantage thermally to introduce 
considerable water along with the air, as has already been pointed out on pages 269 
and 520, as long as the percentage of H in the gas is not high enough to cause 
trouble through pre-ignition. A second reason for limiting the amount of water used 
is found in the producer itself in that, since the decomposition of water requires heat, 
























598 


APPENDIX 


there is a tendency to lower the temperature of the generator, which favors the 
formation of the undesirable carbon dioxide (see p. 600). In general, therefore, we 
find that the higher the percentage of hydrogen in a producer gas, the higher also 
that of carbon dioxide. 

The most efficient hydrogen content is found between 20 and 25%, but it is also 
found that with this percentage only few engines run smoothly. In most engines 
violent explosions and knocking in the crank mechanism usually compel a reduction 
in the water supplied to the generator long before the most efficient percentage of 
H in the gas is reached. The upper limit of water used by the producer is thus 
evidently set in most cases by the amount of hydrogen the engine can stand under 
full load. 

Concerning the quantity of impurities, such as tar, carried by producer gas, a 
point of prime importance in the operation of engines, it is difficult to give any 
numerical data. Only a direct examination of coal or gas can give any reliable 
information regarding the subject. 

The best German hard coals (Rhenish, having on the average 14 400 B.T.U. per 
lb.) according to a number of analyses, form from .12 to .15% of tar, that is from 
.0012 to .0015 lb. per lb. of coal. Assuming that the gas yield per lb. of coal 
averages 75 cu.ft., this would correspond to a tar content of from .016 to .020 lb. 
of tar per 1000 cu.ft. of raw gas. Thorough washing reduces this to from .003 to 
.006 lb. of tar per 1000 cu.ft., and careful wet and dry purification (in sawdust 
purifiers) may bring it down to .0003 lb. or even less per 1000 cu.ft. 

The raw gas from bituminous coals is of course much moie impure, that is, it 
carries more tar. Thus, for instance, a sample of gas (composition in volume-per-cent, 
C0 2 = 1.0, C n H n =.4, O 2 = 0, CO = 31.5, CH 4 = 2.5, H = 11.8, N = 52.8, heating value =176 
B.T.U. per cu.ft.), made from an average sample of soft coal (H~ 12 600 B.T.U.) 
contained in the raw state from 1.27 to 1.45 lbs. of tar per 1000 cu.ft. of gas. After 
a thorough scrubbing in a special washer this was reduced to .006 lb. per 1000 cu.ft., 
which about corresponds to what is found in washed anthracite gas and is yet 
permissible from the standpoint of practical operation. 

6. Thermo-Chemical Diagrams Illustrating the Various Gasification Phenomena. 1 

Figs. 704a-d illustrate the thermo-chemistry of the simple combustion of carbon in air and in water 
vapor. The fundamental equations for the combustion of 1 lb. of C in air are stated as follows: 


C 2 +2 (O +oN) 2 = 2C0 2 + 2aN 2 ,.(1) 

C 2 + (0 + oN) 2 = 2CO + aN 2 ..(2) 


Eq. (1), utilizing the oxygen from 141 cu.ft. of air develops an available amount of heat 
equal to 14 544 B.T.U., while in eq. (2) the oxygen from only 70.5 cu.ft. of air is used, resulting 
in a development of heat equivalent to 4446 B.T.U., the remaining 10 098 B.T.U. being chem¬ 
ically bound in the CO produced. Between these two limits, of course, any combination of these 
two types of combustion is possible. Attention is called at this point to the latent heat of 
carbon which must be expended in converting the solid C to the gaseous form. The fact that 
heat must be so expended is the reason why the heating value of C is not 20 196 B.T.U. per lb., 
but only 14 544, and why the heat developed from the combustion of C to CO is not exactly 
half that of the combustion of C to C0 2 . 


1 Karl Kutzbach. Taken from an article published in the Z. d. V. D. I., 1905, p. 233, with the per¬ 
mission of the author. 





THEORY 


599 


In Fig. 704a, Diagram 1 shows the volume relations between the gases resulting from the 
combustion of unit weight of carbon according to eqs. (1) and (2). This diagram is then 
transposed into Diagram 2, Fig. 7046, showing the volume relations based on the cubic foot, and 
Diagram 3, Fig. 704c, showing the weight relations based on the pound. Diagram 7, Fig. 704d, 
shows the thermal relations existing during both combinations. 

Diagram 1 is obtained as follows: 

Combustion according to eq. (1): 

C 2 + 2 (0 + aN) 2 = 2C0 2 + 2aN 2 .U) 


Weight relation: 24+ 64 + 56a — 88 + 56a 

or 1 lb. C + 2.66 lbs. 0 +2.33a lbs. N = 3.66 lbs. C0 2 +2.33a lbs. N. 

Since the oxygen in this combustion is obtained from air, and the weight relation between N 

and O in a ir=—, the value of a in the above equation must be 
23.5 


76.5 

23^5 


X 2.66 = 2.33a, 
a = 3.73. 


Substituting this value of a in the above equation, we finally ha\e 

1 lb. C + 2.66 lbs. 0 + 8.70 lbs. N = 3.66 lbs. C0 2 +8.70 lbs. N.. (la) 

Now to change eq. (la) into a volume relation we simply employ the specific'height r f or 
each gas from Table 142. The specific weight of gaseous carbon is taken at .0668 lb. pei c . . 

We have for 1 lb. of C, 

Volume Relation : 

14.97 cu.ft.C+29.93cu.ft. 0 + 111.1 cu.ft. N-29.93 cu.ft. CO,+111.1 cu.ft.N. . . (16) 

141 cu.ft of air 

Combustion, according to eq. (2), treated in a similar manner, gives the following relations: 

C, + (0 + aN),-2CO + aN,. 

Weight relation: 

24+ 32 +28a = 56 +28a, 

1 lb. C+1.33 lbs. 0 + 1.17a lbs. N-2.33 lbs. CO+1.17a lbs. N. 

Putting a = 3.73 as before, we ha\e 

1 lb. C + 1.33 lbs. 0 + 4.35 lbs. N-2.33 lbs. CO + 4.35 lbs. N.(2a) 

Volume relation. This is obtained as above. 

14.97 cu.ft.C + 14.97 cu.ft. 0 + 55.55 cu.ft. N- 29.93 cu.ft. CO+55.55 cu.ft. N. . . (26) 


or 


70.5 cu.ft. of air 












600 


APPENDIX 


With the aid of eqs. (16) and (26) the construction of Diagram 1, Fig. 704a, becomes obvious. 
The ordinate marked 2 (in a circle) represents the relation of volumes at the end of the 
combustion according to eq. (2), that is, after 70.5 cu.ft. of air have been supplied. The original 

volume of 1 lb. of gaseous C, 14.97 cu.ft., has 
at this point been reduced to zero, and 29.93 
cu.ft. of CO have appeared in its stead. At 
the same time 55.55 cu.ft. of N were supplied 
along with the O, making the total cubic feet 
of gas 85.48. Now as the supply of air is 
increased beyond 70.5 cu.ft., C0 2 commences 
to make its appearance at the expense of CO, 
until, when the supply of air has reached 
141.03 cu.ft., all of the CO has disappeared, 
while 29.93 cu.ft. of C0 2 appear in its place. 
The cubic feet of N supplied in the meantime 
has increased to 111.10, making the total 
volume of gas along the ordinate marked 1 
(in circle), indicating the end of combustion 
according to eq. (1), equal to 141.03 cu.ft. 
For any air supply beween 70.5 and 141.03, 
both CO and CO, appear in the diagram, the 
relation between them always being 


Volume CO + volume C0 2 = 29.93 cu.ft. 


The conventions used in this diagram, 
together with some other data relating to 
some of the other diagrams, are shown in 
Fig. 703. 

Diagram 2. Fig. 7046, is obtained from Dia¬ 
gram 1 by merely expressing the volume relations 
existing at ordinates (1) and (2) in percentages 
of the total volume, while Diagram 3, Fig. 704c, 
is constructed by aid of eqs. (la) and (2a). In Diagram 2 the curve of heating value has been 
obtained by determining the heating value of .347 cu. ft. of CO. This is equal to 117 B.T.U. Now 
as more air is supplied some of the CO burns to C0 2 , and the heating value of each cubic foot 
of the gas is less as we go beyond 70.5 cu.ft. of air, until at ordinate (1) all of the CO has 
been burned to C0 2 , and the heating value is then reduced to zero, all of the gas being incom¬ 
bustible. The curve is plotted to the same scale as the scale of percentage at the side of the diagram. 

If steam be led through incandescent carbon, the resulting gas consists of a mixture of H, 
CO, and C0 2 . The reaction proceeds according to either or both of the following equations: 


C 2 + 2H 2 0 = 2CO +2H,,.(3) 

C 2 + 4H,0 = 2CO, + 4 H,.( 4 ) 


Determining the volume and weight relations represented by these equations, as before, we 

have 

From eq. (3): 

Weight relation: 24 + 36 = 56 + 4, 


Convention 

Used. 

Wt. in Lbs. 
perCu. Ft. 

Heating 
Valve per 
Cu. Ft. 

Cu. Ft. 
per Lb. 

Heating 
Value per 
Lb. 

ra 

— 

— 

14.97 

14500 

tMlIl 

.08921 

— 

10.22 


.07807 

342 

12.81 

4320 

cqwm 

.12267 

— 

8.17 

— 

n m 

.07831 

— 

12.81 

— 

Ur e/Vot % 

Ljm 

.08072 

— 

12.51 

— 

U ra 

.00559 

Higher 

346 

297 

Lower 

179.7 

Higher 

61560 

51730 

Lower 

Hj,oV' s 1 

.05016 

— 

19.4 

— 


Fig. 703.—Key to Figs. 704 and 705. 


or 


1 lb. C + 1.5 lbs. H,0 = 2.33 lbs. CO + .166 lb. H 


(3a) 


This equation is converted to the volume relation 
Table 146. 


by the aid of the specific weight y from 























































THEORY 


601 


Volume relation: 14.97 cu.ft. of C + 29.93 cu.ft. H,0 = 29.93 cu.ft. CO + 29.93 cu.ft. H. . (36) 
From eq. (4): 

Weight relation: 24 + 72 = 88 + 8, 

1 lb. C + 3 lbs. H 2 0 = 3.66 lbs. C0 2 + .333 lb. H.(4a) 

For this we obtain, as before, 

Volume relation: 14.97 cu.ft. C + 59.88 cu.ft. FI,0 = 29.93 cu.ft. C0 2 + 59.88 cu.ft. H. . . . (46) 


Equations (3a), (36), (4a), and (46) serve to plot Diagrams 4, 5, and 6, Figs. 704 a to c. 

The method of doing this is similar to that used for Diagrams 1, 2, and 3, and need hardly be 

outlined any further. The heating value curve in Diagram 5, Fig. 7046, is obtained by deter¬ 
mining the heating value of .667 cu.ft. of H at ordinate (4), and the heating value of .5 cu.ft. of 
CO+ .5 cu.ft. of H at ordinate (3). This determines the end-point of a curve for which inter¬ 
mediate points may be found from the volume relations in the same diagram. 

Finally, Diagrams 7 and 8 in Fig. 704d show the various quantities of heat concerned in the 
reactions for both air gas and water gas. The construction of these diagrams is obvious, the 
various quantities of heat being computed directly from the weights involved in eqs. (la), (2a), 
(3a), and (4a). 

All of the diagrams between the end conditions at ordinates (1) and (2), or (3) and (4), show 
interactions resulting in the formation of both CO and C0 2 , and the question next comes up as to 

what influence determines the relative formation of these two gases. This controlling factor is 

the temperature. There is a chemical law which states that with increasing temperature the 
chemical actions and interactions are so modified as to resist the further increase of temperature, 
that is, that those reactions which take place with a smaller development of heat or which are 
combined with a cooling process gain the ascendency over other possible reactions. In the air 
gas process, the reaction combined with the smaller development of heat is evidently that accord¬ 
ing to eq. (2), that is, the formation of CO. In the water-gas process the reaction showing the greatest 
cooling effect is that according to eq. (3), again the formation of CO, because as Diagram 8, 
Fig. 704d, shows, at ordinate (3) the latent heat in the gases is 18 679 B.T.U., while at ordinate 

(4) it is only 17 244 B.T.U. , , 

In actual practice the temperature is of course by no means the same m all parts ol the 
producer Under normal conditions, it probably increases from the grate upward until all of the 
oxygen in the air has been combined, after which it decreases on account of the heat furnished 
to the green coal moving downward. In any part of the producer the relative formation of CO 
and C0 2 depends entirely upon the temperature, the transformation from the C0 2 formed to CO 
in the hotter zones going on at the same time that some CO is again converted into CO, in the 
cooler zones. The composition of the producer gas as it finally leaves the generator is a com¬ 
promise between these combinations and dissociations at different temperatures, but the net result 
is that the highest temperature, on account of its greater potency, is the deciding factor. 

Examining again the heat diagrams for air gas and water gas, it will be noticed that tie 
former shows a considerable quantity of sensible heat, which as far as the obtaining o le 
greatest possible amount of chemical energy is concerned, is wasted. The heat diagram for the 
water-gas 1 process on the other hand shows a deficiency of heat, since C itself can only furnish 
14 544 BTU per lb. At ordinate (3) there is a deficiency of 4176 B.T.U.; at ordinate (4) of -/00 
BTU Both quantities referred to steam at 32° F. If now both steam and air are introduced 
together, the sensible (waste) heat of the air-gas process can supply or balance the deficiency 

0t th ThrSams r o7 Figs. 705a to c illustrate this combined air-steam (producer) gas process. 
Diagrams 9 10 aud 11 refer to the perfect gasification process, the efficiency being assumed 100%. 

%iagram 9 Fig 705a, showing the “heat balance” for the perfect process, indicates the method 
Of findffigThe balance or state of equilibrium above mentioned. The three-coordinate system is 
used the three coordinates representing the variables: air, steam, and heat units as indicated. 
Tim ' diagranihi the left vertical plane is a reproduction of Diagram 7 Fig. 704d; that m the 
right veS p ane one of Diagram 8. With their aid it is possible to determine the conditions 
It Which the waste heat of the air-gas process just balances the deficiency of the water-gas 



Air Gas Process. Water Gas Process. 























































































































































Water Gas Process. 


o03 


\ 





Fig. 704b.—Volume Diagrams for One Cubic Foot of Gas, Air Gas and Producer Gas Processes. 








































































































































































Air Gas Process. Water Gas Process. 


604 


O 



CO 


Fig. 704c— Weight Diagrams for Air and Water Gas Processes. 















































































































Water Gas Process. 


00 


505 





■ * 1 ■ i _j_i—i—i—i—i—i—■—»—*—>—■—*— 1 —■—■— '- 



p 

£ 


c3 

a 


c3 

£ 

-a 

c 

ci 


be 

S3 


TT 

O 

o 

£ 


























































































































Efficiency 100% 



I ig. 705 a. Heat Balance Diagrams for Producer Has Process. 

































































































































♦ 


Fig. 705 b .—Volume Diagrams per Pound of Carbon Gasified, Producer Gas Process. 




































































































































Efficiency 100% Efficiency 75% 

3/3B ruk 3/3 B Tu 






Fig. 705c.—Volume Diagrams for One Cubic Foot of Gas, Producer Gas Process. 













































































































































THEORY 


609 


process. Points c, d, e, and / defining the waste heat of the air-gas process, are connected as 
shown with points k, i, h, and g, defining the heat deficiency of the water-gas process. The points 
of intersection a and 5 and their vertical projection a { and 5 U defining the producer-gas diagram, 
allow of the determination, in the horizontal plane, of the relative quantities of air and water. 

The diagram shows that at the point a t for instance, the quantities to maintain a balance 
furnished per pound of C should be .77 lb. of steam and 34.5 cu.ft. of ay. At this point there 
will be no C0 2 found in the gas. At the point 6, the respective quantities per pound of C are 
22.46 cu.ft. of air and 2.53 lbs. of steam. At this point all of the CO will have been converted 
to C0 2 . 

The upper-end point 0 \ of the line defining the heat in CO in the producer-gas diagram is 
found by drawing the line of. The CO area in the producer-gas diagram is, in the nature of the 
case, the same as in the other two diagrams. The total sensible heat of the air-gas diagram 
now appears as heat in H in the producer-gas diagram. 

The assumption of a generator efficiency equal to 100% presupposes that the sensible heat 
of the materials entering the gasification process is at least equal to the sensible heat of the 
gas leaving the producer. In actual practice, however, the gases leave the gasification zone of 
the producer at the combustion or reduction temperature. They consequently carry along a 
considerable stock of sensible heat of which only a small part is transferred to the coal on the 
way out for drying, distilling, etc., and another part is recovered for pre-heating purposes. A 
large part of this sensible heat is consequently lost, and the efficiency of the process is therefore 
at once reduced, the actual gasification efficiency ranging from 75 to possibly 85% as a 
maximum. 

If we assume as a lower limit an efficiency of 75%, the available heat is reduced from 
14 544 to 10 908 B.T.U. per lb. of C, while the waste heat of the air-gas process is only 810 
B.T.U. This of course means that less heat is available for the water-gas process and that 
consequently less water can be evaporated. The heat balance for this process is shown in 
Diagram 12, Fig. 705a, the construction of which is obvious. The producer-gas diagram in this 
case shows, when no C0 2 is produced (point a), that to maintain a balance .14 lb. of water and 
64.18 cu.ft. of air must be furnished per pound of C. 

Diagrams 10 and 13, Fig. 7055, next show the volumes of the various gases produced per pound 
of carbon gasified. Their construction, once the position of what might be called the “equilibrium” 
line af > t is found, is easy and obvious. 

Finally, Diagrams 11 and 14, Fig. 705c, expressing the composition of the producer gas made in 
volume-percent are directly obtained from Diagrams 10 and 13. These diagrams also show the 
curve of heating value per cubic foot of producer gas. Attention is called to the decided 
variation in the production of hydrogen depending upon the producer efficiency and the amount 
of C0 2 made. 

Up to this point all of the discussion of the gasification process has been based upon the 
assumption that pure carbon was used. In practice, however, we find carbon available only with 
certain admixtures. Coke always contains less carbon than the undistilled fuel because a part 
of the carbon together with hydrogen forms the volatile ingredients, like tar and the light 
hydrocarbons of illuminating gas. Distillation commences at comparatively low temperatures, 
900 to 1400° F., and is therefore likely to be complete when the gasification zone of the producer 
is reached, especially for small sized fuel. The only important factor in the gasification process 
itself is then the carbon in the coke. To obtain the composition of the gas resulting in such 
a case, the gases of distillation (illuminating gas) obtained for 1 lb. of C may, according to volume 
and composition, be drawn in above the quantities of producer gas shown in the volume diagrams 
(either 10 or 13) of Fig. 7055. Under certain circumstances this may result in a considerable 
increase in the heating value of the producer gas, since the heating value of illuminating gas 
may be from 450 to 575 B.T.U. per cu.ft. The percentage of CH 4 found in producer gas and a 
part of the H content is always due to the gases of distillation from the coal as fired. The 
total efficiency of the producer consequently always exceeds the mere gasification efficiency, and 
the gain is the greater the higher the percentage of volatile matter (hydrocarbons) in the coal. 

In Diagram 14, Fig. 705c, the range of producer gases usually made in the normal operation 
of a producer plant, depending upon the supply of steam, is indicated by the area in the producer- 
gas diagram marked by marginal hatching. Of these gases the one containing most H, and 
consequently also most C0 2 , corresponds about to the gas made by the Mond process, although 
the H content of this gas may, with coal high in volatile matter, be even considerably higher 
than that shown. The low-temperature operation of the Mond process permits ot the obtaining 


610 


APPENDIX 


of an excellent producer efficiency, which is, however, somewhat reduced owing to the large 
quantity of steam (up to three-fourths of that furnished) passing through the generator 
undecomposed. 


B. SOME DETAILS FROM PRACTICE 1 

I. Directions for Operation, Attendance, Etc. 

Instructions for the operation and care of internal-combustion engines, intended for 
the layman, as distinguished from the engineer or skilled attendant, can hardly be 
made too elementary. Such directions should clearly and concisely point out the 
operations that must be gone through in daily routine and should, besides this, also 
give quick and safe advice in the case of any troubles out of the ordinary that may 
arise. An expert erector will personally give such instructions to the regular operator 
in a very few hours, but the written word, as always, calls for distinctness and 
exactitude even at the expense of brevity. 

There is of course little use in presenting here any sets of instructions as examples, 
because each different make of engine requires its own distinct set. But in order to 
show both the language employed and the very great help given by pictures, etc., 
of certain parts, there are given below instruction books and directions issued by one 
or two well-known American firms. 

Besides giving a- general key to the main engine parts, any set of instructions 
must consist of comprehensive and logically arranged series of rules for starting and 
stopping and should include hints for the hunting down of troubles and for eradicating 
them. 

Instructions concerning producer installations in general read much alike, no matter 
what the make of the producer. The reason for this is that, although the various 
operations required depend somewhat on the constructive details of the producer, the 
latter are much less diverse than is the case in gas engines, and, fundamentally 
considered, all gas producers are nothing but a special type of furnace. Directions 
for judging of the proper state of the gasification process, for recognizing and removing 
unusual troubles, etc., are consequently much the same in all cases. The following 
pages give the directions and instructions issued in the case of three well-known 
makes of producer installations, mainly to serve as a guide for similar purposes. 

Special emphasis should be laid on such statements or points in the instructions, 
the neglect of which may lead to dangerous occurrences (accidents to attendants or 
to machinery). In this connection it should be remembered that in some instances 
of this kind the manufacturers may be legally responsible to either the purchaser or 


1 Translator’s Note. This division of the book does not strictly follow the German text, for the 
reason that some of the material there given is of no importance to American readers. Thus the directions 
given for the operation and care of a kerosene engine have been replaced by others issued by American 
makers. The instructions concerning gas producers were translated for reasons given below. 

The original text contains nothing concerning the methods of testing internal-combustion engines, 
and on account of the importance of this subject both the Code of the American Soc. of Mech. Engrs. and 
that of the German Society of Engineers are inserted. It is understood that the former is at the*present 
writing being revised by the Gas Power Section of the American Society.. 




SOME DETAILS FROM PRACTICE 


Oil 


the persons injured, and for that reason any instructions aiming to point out the 
possible serious consequences of certain wrong management or attendance should be 
made specially conspicuous. 

1. Instruction Book for Olds Kerosene Engines, Type AK 


Olds Gas Power Co., Lansing, Mich. 

DON’T WORRY 

Do not let the size of this book alarm you. O’. .’..3 engines are simple, easy to run, easy to 
understand and reliable. We have made this book large simply because we wish to cover every 
possible point and put you in shape to take care of your engine yourself. If you study this 
book and understand it, you will have no trouble. If you do have trouble, whatever it is, if 
you read this book and understand it you can fix it yourself. 

OLDS GAS POWER CO 

LANSING - - MICHIGAN.U.S.A. 

TYPE AK ENGINES 

C 

-W- 


v/ 


I 


~M 


ik: 




W 3 - 


L 


Xz~ 


o 



L-571. 


Type 

T 

k 

A: 

Dimensions Given In Inches. 

Pull 

EYS 

General Dimensions. 

Engine Base 

Sub Base 

AnchorBoli 

Wheel 

5t 

D. 

5ma 

'LL 

A 

F 

r. 

n 

F 

F 

6 

H 

/ 

J 

K 

L 

M 

N 

0 

F 

h 

a 

R 

S 

T 

U 

1/ 

W 

X 

y 

Z 

a 

b 

Via. 

BELT 

DlA. 

Ear 

m 

,? 

5% 

15k 

74? 

401 

?il 

iti 

bi 

li 

111 

Z 

li 

51 

77 

II 

3 

70i 

9i 

z 

8i 

78 

14 

8k 

8 

7bi 

I7z 

2 

z 

16 

78 

Z 

6 

4 

— 

— 

w 

4? 

w 

17? 

Ill 

44\ 

743 

Ili 

Si 

ti 

73 

n 

H 

bl 

75i 

13 

5k 

74 

III 

1 

2 

9l 

37 

16k 

9 

81 

TOi 

15 

1 

2 

z 

lb 

Z8 

li 

8 

5 

b 

4 

m 

ft 

490 

14 


4bl 

?bi 

ml 

91 

li 

M 

3i 

z 

7i 

77 

14 

6i 

75i 

izi 

2 

\ol 

34 

I8i 

to 

U 

3?i 

16} 

5 

3 

7i 

70 

33 

n 

10 

5 

— 

— 

WK 

ft 

479 

?(l? 


sol 

111 

ml 

It 

u 

79% 

4i 

z 

bl 

28} 

I4i 

6 

77 

li 

A 

6 

as 

13 

36i 

Mi 

I8i 

II 

9s 

34i 

17 

£ 

3 

li 

70 

3b 

3 

IZ 

b 

— 

— 


IZ 

m 

Hi 

41 z 

% 

m 

I7i 

m£ 

Hi 

M 

4& 

It 

9 

35i 

18 

8 

$ 

lb 

J 

A 

73 

4 

IZ 

4/2 

7/ 

3 

A 

7i 

78 

40 

3i 

18 

6 

— 

— 


Fig. 706. 


Four-Cycle. The Olds engine belongs to the “ four-cycle” type-that is, it gives one explosion 
every two revolutions of the fly-wheel. Each two revolutions of the fly-wheel complete a cjcle 
of four operations. Look at your engine, and we will take them in their order 

1 Charaina or Suction Stroke. The crank has pushed the piston as far back into the 
cylinder as it will go, and now as it goes on turning, it starts to pull the piston out again - 
the inlet valve opens and as the piston moves forward, the suction pulls in a chaige of mixet 
fuel and air until the space behind the piston is completely filled, when the inlet valve closes. 

2 Compression Stroke. The piston has gone out its full distance, the cylinder is full of gas, 
and the pTston now starts back in, compressing this mixture of gas and air. Just as the piston 
is reaching the end of its compression stroke, the electric spark ignites the charge, t e 































































































































































612 


APPENDIX 


3. Explosion or Expansion stroke takes place, and the piston is pushed out with great force 
by the explosion of the gas. This is the working stroke, the energy of which is transmitted to 
and stored by the fly-wheels. The piston rushes out to its full limit and returns on the 

4. Exhaust Stroke. Before the working stroke was quite completed, the exhaust valve had 
been already thrown open to avoid back pressure, and the returning piston now simply sweeps 
the cylinder of burnt gases, driving them before it and out through the exhaust valve. Your 
cylinder is cleaned and swept, so to speak, ready for a new charge, and your cycle is complete. 

INSTRUCTION BOOK OLDS TYPE AK ENGINES 

This booklet is intended to instruct explicitly in the operation of the Olds kerosene engine. 
It has been gone into so carefully and so thoroughly that if it is studied intelligently it will be 
easy for anyone to run our engines and keep them running, whether experienced or not. 

Before the engine leaves our hands the adjustments have all been made by us, and it is 
ready to run. 

Do not tamper with anything until you have fully mastered these instructions, and know 
just what you are doing, and why, and what the result will be. Do not turn a nut to tighten 
it or loosen it, or fix something that may seem loose, or loosen something that may seem 
tight, just out of curiosity, as this will generally make trouble for you. 

Remember that it is absolutely necessary to use gas engine oil in lubricating cylinder and 
piston—no other kind of oil will do. 

Instructions for Setting Up. Do not remove the crating until the engine has been brought 
to the place where it is to be used. In taking off the crating, be careful not to break anything 
inside. Set the engine on the foundation, and always remember that the firmer the foundation 
the smaller the repair bills. If a cement foundation is used there must be anchor bolts coming 
up through the cement base to fit the holes in the bed of the engine. If the engine is placed 
on the floor or on skids use lag bolts or regular bolts. 

The table of dimensions (Fig. 706) gives the exact size of these, and the location of these 
holes. 



Ignition System. The ignition system is what is known as the jump spark, in which an 
induced current of high tension is caused to jump a gap between electrodes in a plug screwed 
in the cylinder head. The current for this system can be supplied by batteries, magneto, or 
both together. Below are shown three cuts. Fig. 707 shows the connections between the 
batteries, the coil, the contact maker, and the spark plug. For clearness the cut shows the 
batteries and plug enlarged out of proportion to the engine. 

Fig. 70S shows the arrangement of the batteries and magneto where these two systems 
are installed together, either one to be used. 

Fig. 709 shows arrangement where magneto is used alone. 



































SOME DETAILS FROM PRACTICE 


613 


Using Battery Only. In the crate you will find a large red box. This is the battery box 
The spark coil and the batteries are inside—and they are properly wired together—on the 
inside. See that in each case the connection is tight—screwing down any of the post screws 
that happen to be loose (one on the coil and two on each battery at the point where the wires 
make the contact). 

The wires come through a hole in the battery box, and on the outside of this battery box 
is a coil of wire. This wire is to be used in connecting up the battery box with the engine. 

Open the battery box with a screw driver, and you will find tw r o spark plugs in the box; 
one is for actual use, and the other is for reserve. 

See that the points in the spark plug are about ^ of an inch apart. If they are too far 
apart the engine will skip, and if they are too near together you will not get a big enough 
spark to explode the charge. You can bend them slightly with a pair of pliers. Be very 
careful in doing this, as the spark plug is delicate. 

Set the battery box near enough to the engine to have the wires reach easily. The large 
wire is to be fastened to the spark plug E, Fig. 707. If there is a terminal attached to the 
wire, this fits under the screw on the spark plug. Force it gently around the spark plug post 
and turn the thumb screw r down until it holds tightly. If there is no terminal on the wire, 
cut aw r ay some of the insulation, or rubber covering, at the end of the v r ire, and take this bare 
wire and wrap it around the top of the spark plug post, then tighten it with pliers so that the 
connection is tight. Take the small wire that comes out of the box nearest to the big wire 
and fasten this on top of the commutator D, Fig. 707. The other wire coming out of the box 
nearest the switch is fastened to the engine bed under the round-headed screw G, Fig. 707. 
See that the short insulated wires connecting the five cells together are screwed down tight. 
Should any wires become disconnected, see diagram, Fig. 707, for wiring. 



Using Batteries and Magneto. Batteries and magneto are to be arranged so that they can 
be run alternately at will. The engine should be started on the batteries, the batteries then 
thrown off and the running done on the magneto. In Fig. 708, H is the magneto, and F is 
the two-point switch. When the switch is on point J (running position) the magneto is 
connected in the circuit, and furnishes current to the system; and when it is on point A 
(starting position) the battery furnishes the current. 

The arrangement (Fig. 709) shows the connections where the magneto alone is used. 

If a magneto is used, the commutator should be kept clean with a little gasoline and waste, 
and any glaze which is on the rubbing surface of the brushes should be removed with a knife. 
The brushes are purposely made difficult to get at, and without adjustment, as once set they 
should remain in adjustment until worn out. A direction sheet accompanies each magneto, 
and goes into the subject of the magneto quite thoroughly. The magneto furnishes a direct 













































614 


APPENDIX 


current of low tension of from 7 to 9 volts. If it is desired to substitute batteries for magneto, 
or supplement the magneto with the batteries, enough batteries should be supplied to give 
6 volts. 



Fig. 709 


The Coil. The spark coil C, shown in Figs. 707, 708, and 709, consists of a primary coil 
through which the current from the batteries or the magneto is carried, the vibrator which 
interrupts the circuit automatically in this primary coil, and the secondary coil in which is 
induced a current of high voltage, due to the interruption by the vibrator of the current in 
the primary coil. ! urther, in the primary circuit, is interposed a condenser made of tin foil 
and paraffin paper. 

The principal point to avoid in the use of coil is dampness; if the coil becomes soaked 
with rain, it should be dried out very slowly at a low temperature, so as not to melt the 
paraffin in the condenser. 

Muffler and Water Piping. Fig. 710 shows the muffler piping for all sizes; Fig. 711 shows 
the water piping for 2AK and 3AK engines. Fig. 712 shows same for 4AK, 5AK and OAK. 
Below each cut is a little printed table giving the location and size of each pipe and fitting 
for each engine. The piping and fittings will be found in a box in the crate of the engine; 
and if careful attention is given to putting on the connections exactly as shown, they will be 
found to go together properly. If you have to carry the exhaust further from the muffler 
than the short pipe furnished will reach, the extra piping for this purpose must be of a larger 
size than the pipe from the muffler. In piping the exhaust outside a building, have the hole 
in the building through which it goes out lined with tin, or some non-conductor of heat to 
prevent the wood of the building being scorched. Always have the muffler as close to the 
engine as possible, and carried on foundation or support of some kind so the weight of the 
muffler will not be hanging on the flange of the valve cage. In case it is desirable to pipe 
the exhaust out doors, it is usually advisable to set the muffler just outside the building, 
provided this can be done by carrying the muffler not more than 10 ft. from the engine. The 
muffler should be placed in such a position that it cannot fill up with water, either from rain 

oi any other cause. Long exhaust pipes are bad. They cause back pressure on the engine; 

turns in the pipes should be avoided and the exhaust should be allowed to go as directly out 
of the muffler as possible. If the exhaust pipe is long, condensation of moisture might drain 

back into the muffler and seal the muffler off, preventing starting. There is a drain on the 

bottom of the muffler. Open this drain occasionally and drain off any water in the muffler. 















































SOME DETAILS FROM PRACTICE 


615 


In the water pipe system drains are provided for the tank and cylinder jacket. It is 
advisable to clean out the tank and drain off the water from the jacket occasionally to remove 
any dirt or scale which may accumulate; and in cold weather the water should always be 
drained off to prevent freezing and breakage. 



EXHAUST PIPING 


2AK. 1. HX12" W. I. Pipe. 

2. 11X I" Reducing Tee. 

3. 11" Close Nipple. 

4. ll"X6'-0" W. I. Pipe (thread one end 

only). 

5. 1X81" W. I. Pipe. 

6. 1" Pipe Plug. 

7. 1" Mall. Tee. 

8. HXli" C. R. S. with Pipe (thread 

on both ends). 

9. 1X4" C. I. Flange. 


3AK. 1. 11X12" W. I. Pipe. 

2. 11X1X11" Reducing Tee. 

3. 11" Close Nipple. 

4. ll"X6'-0" W. I. Pipe (thread one end 

only). 

5. 1X101" W. I. Pipe. 

6. 1" Pipe Plug. 

7. 1" Mall. Tee. 

8. ifXll" C. R. S. Pipe (thread on 

both ends). 

9. 1X4" Flange C. I. 


EXHAUST 

4AK. 1 . 2X12" W. I. Pipe 

2. 2X1X2" Reducing Tee. 

3. 2" Close Nipple. 

4. 2"X6'-0" W. I. Pipe (thread one end 

only). 

5. IX 10f" W. I. Pipe. 

6. 1" Pipe Plug. 

7. 1" Mall. Tee. 

8. RX11" C. R. S. Pipe (thread on 

both ends). 

9. 1X4" C. I. Flange. 

5AK. 1. 2X12" W. I. Pipe. 

2. 2X1X2" Reducing Tee. 

3. 2" Close Nipple. 

4. 2" X 6'-0" W. I. Pipe, (thread one end 

only). 

5. 1X111" W. I. Pipe. 


PIPING 

6. 1" Pipe Plug. 

7. 1" Mall. Tee. 

8. HXli" C. R. S. Pipe (thread on 

both ends). 

9. 1X4" C. I. Flange. 

6AK. 1. 21X14" W. I. Pipe. 

2. 21X11X21" Reducing Tee. 

3. 21" Close Nipple. 

4. 2l"X6'-0" W. I. Pipe (thread one end 

only). 

5. 11X10!" W. I. Pipe. 

6. 11" Pipe Plug. 

7. 11" Mall. Tee. 

8. 11X 2" C. R. S. Pipe (thread on both 

ends). 

9. 11X5" C. I Flange. 


In case running water is used instead of a tank the doming j^ketw^rs^ui d tetaken in at 

the lower opening of the jacket and the over ®°^ * “ ° f U t £ e tank to the under side of the jacket, 
is used, the circulation of water is from the bottom of the tankm^ ^ ^ ^ ^ ^ 

up through the jacket out through the p £ in the tank is always above the 

FiU the water tank w.th clean in the morning, and keep the level 

return pipe, which is the top one. Be P 0therwise n0 c i rcu lation will occur around 

in the tank always over the outlet of the top P P' • AlgQ be gure the leyel G f t he water in 
the cylinder, and the cylinder will become • bottom of the tank is on 

the tank is above the cylinder. The‘ ^ t^the tank from the top of the engine 

^ "^vei* of Z^ tank will be above the cylinder. 















616 


APPENDIX 




COOLING WATER PIPING 


2AK. 11. ljx 18" Rubber Hose. 

12. 1X3" W. I. Pipe. 

13. 1" Mall. Tee. 

14. 1X12" W. I. Pipe (thread one end 

only). 

15. 1" Lock Nuts. 

16. 1|X2" Rubber Washers. 

17. 1" Mall. Tee. 

18. 1" Pipe Plug. 

19. 1X4" W. I. Pipe (thread one end 

only). 

20. 1X31" W. I. Pipe. 

21. 1" Brass Cocks. 

23. 1X3" W. I. Pipe. 


3AK. 11. If X IS" Rubber Hose. 

12. 1JX8" W. I. Pipe. 

13. 11" Mall. Tee. 

14. 11X12" W. I. Pipe (thread one end 

only). 

15. 11" Lock Nuts. 

16. lfX2f" Rubber Washers. 

17. 11" Mall. Tee. 

18. 11" Pipe Plug. 

19. 11X4" W. I. Pipe (thread one end 

only). 

20. 11X4" W. I. Pipe. 

21. 11" Brass Cocks. 

23. 11X4" W. I. Pipe. 


COOLING WATER PIPING 


4AK. 10. 11X12" Rubber Hose. 

11. lfX18" Rubber Hose. 

12. 11X6" W. I. Pipe. 

13. 11" Mall. Tee. 

14. 11X12" W. I. Pipe (thread one end 

only). 

15. 11" Lock Nuts. 

16. llX2f" Rubber Washers. 

17. 11" Mall. Tee. 

18. 11" Pipe Plug. 

19. 11" Close Nipple. 

20. 11X12" W. I. Pipe (thread one end 

only). 

21. 11" Brass Cocks. 

22. 11" X 4" W. I. Pipe (thread one end 

only). 

23. 11X31" W. I. Pipe. 


5AK. 10. 11X12" Rubber Hose. 

11. lfX18" Rubber Hose. 

12. 11X10" W. I. Pipe. 

13. 11" Mall. Tee. 

14. 11X12" W. I. Pipe (thread one end 

only). 

15. 11" Lock Nuts. 

16. If X2f" Rubber Washers. 

17. 11" Mall. Tee. 

18. 11" Pipe Plug. 

19. 11" Close Nipple. 

20. 11X12" W. I. Pipe (thread one end 

only). 

21. 11" Brass Cocks. 

22. 11"X4" W. I. Pipe (thread one end 

only). 

23. 11X31" W. I. Pipe. 


6AIv. 10. 21X12" Rubber Hose. 

11. 21X18" Rubber Hose. 

12. 2X81" W. I. Pipe. 

13. 2" Mall. Tee. 

14. 2X12" W. I. Pipe (thread one end 

only). 

15. 2" Lock Nuts. 

16. 2f" Rubber Washers. 

17. 2" Mall. Tee. 


PIPING 

18. 2" Pipe Plug. 

19. 2" Close Nipple. 

20. 2X15" W. I. Pipe (thread one end 
only). 

21. 2" Brass Cock. 

22. 2X41" W. I. Pipe (thread one end 
only). 

23. 2X4" W. I. Pipe. 


COOLING WATER 

6AK. 























SOME DETAILS FROM PRACTICE 


617 


Oil Cups. Take the large sight feed oiler, and screw it into a hole on top of the engine 
at H, Fig. 713, and fill this cup with gas engine cylinder oil through the little screw cap at 
the side of the top. Be sure to screw back the cap so that the oil will not splash out. You 

will notice underneath this oil cup a little opening covered with glass, and when the oil is 

running you can see the oil drop down. There is a double thumb screw which regulates these 
drops very easily, as you can see by trying it. 

See that the cup is adjusted to drop about 15 drops a minute for the first two or three 
days your engine runs, then decrease the supply to eight or ten drops a minute, which is about 
what you should use steadily. 

Fill the grease cup on the connecting rod with hard oil. Be sure the cup is kept filled 
with grease, as it lubricates the pin-bearing of the crank-shaft. 

To oil the main bearings you will find oil recesses on top of the main shaft-bearing caps. 

These are filled with cotton waste soaked in oil, and this waste belongs in there. Do not take 

it out, except to replace it with fresh waste. Keep this waste thoroughly saturated with good 
lubricating oil. 

Be sure to oil the rocker arm at L that operates the valves—it is acted on by the push 
rod, and should run freely. (Fig. 713.) 



To put on the pulley, fit it on the wheel opposite to the wheel which carries the governor. 
Three bolts are furnished, and are on the pulley, together with the nuts to hold them on the 

Filling with Kerosene. The engine comes all piped up for kerosene. The kerosene is 
contained in a compartment in the base of the engine. To fill with kerosene, remove the p ug 
at K Fig. 713, place a funnel in the hole, and pour in kerosene, until compartment is filled 
to within two inches of the top. Put the plug back tightly. Notice that a similar opening 
is on the opposite side of the engine for water, and the plug for this one marked is 11, ig. • • 
The kerosene must be put in the opening at K. Be sure that no mistake » made and the 
kerosene put in the wrong side of the engine. The word “Kerosene, is cast in the side of 
the sub-base, in which side the kerosene should be put. 



















































































618 


APPENDIX 


To see whether there is sufficient kerosene in the base put a clean rod down through the 

hole K, Fig. 715. If the rod shows more than in. of its length wet there is enough kerosene 

to run the engine; 3^ in. must show, however, before the engine will run. This is because the 
rod goes down into a small well in the compartment, from which well the kerosene is drawn 
through the suction pipe. The well is below the level of the bottom of the kerosene compart¬ 
ment and the compartment may be entirely dry while the well shows 31 in. of kerosene. 

The same holds good of the water compartment. 

At the opposite side of the engine, in Fig. 715, is the other opening, M, mentioned above. 
This is to be filled with clean water. Both water and kerosene should be strained through a 
fine mesh of wire gauze before being put into the engine. The water and the kerosene are 
both used to make the mixture for the engine cylinder, and the reason for straining both is to 
prevent any dirt from getting in to clog the adjusting valves. The water is fed to the engine 
with the kerosene to prevent pre-ignition in the cylinder. If the engine knocks or pounds when 
the connecting rods and bearings are properly adjusted, a slight admission of water will take 

the pounding entirely away and the engine will run with great smoothness. It will also reduce 

the fuel consumption of the engine. 

The mixer, or carburetor, N, Fig. 715, is shown in section in Fig. 716; that is, it is cut 
through and the drawing is shown as looking into one-half of it. In the section shown in Fig. 
'16, two reservoirs, A and B, are shown as being on opposite sides of the central tube C. In 
the actual mixer, they are both on the same side of the central tube, and are only shown 
opposite in the drawing for greater clearness. The use of these reservoirs is to contain a supply 
of kerosene and water at the level DE shown. Every time there is a suction stroke the engine 



Fig. 714. 


draws in a charge of air through the opening F, which is an open passage from the air into 
the central tube of the mixer, a slight vacuum being formed above the narrow neck of the 
central tube G. The operation of drawing the kerosene and water up into the reservoirs and 
discharging the excess back into the tank below, is similar in both the kerosene reservoir and 
in the water reservoir. If we describe the kerosene, you will understand both. The vacuum 
mentioned above, acting through the port H, Fig. 716, draws the kerosene up from the 
compartment in the base; a small spurt comes up through the pipe K every suction stroke. 
This small spurt amounts to a few drops, and does not fill the reservoir L, but as soon as the 
suction stroke is completed the air rushes in through the port F into the central tube, and 

during the remaining three strokes of the engine atmospheric pressure is maintained in the 

central tube and the reservoir L and the few drops which came into the reservoir L during 
the suction stroke, drop down into tube M, past valve N into reservoir B\ and after this 

has occurred a few times the level of kerosene in the reservoir B will rise until it begins to 

flow out of pipe 0. Any more kerosene coming down from the tube M will simply flow out 
of pipe 0 back into the base of the engine. In this way a constant level is maintained in 
compartment B, which level is slightly below the top of the nozzle P. This nozzle is connected 
by ports with compartment B, and has needle valve Q controlling the flow of kerosene from 
compartment B to the nozzle. 
















































































SOME DETAILS FROM PRACTICE 


619 


As the air rushes past the nozzle in the throat G, it sucks this kerosene up through the 
nozzle with it, making a fine spray. This spray is controlled by the needle valve so that the 
proportion of air and kerosene are exactly right to cause good combustion in the cylinder. The 
spray is thus sucked into the cylinder on each suction stroke, and is then compressed, fired, 
and discharged through the exhaust. 

The water comes through the reservoir A, through needle valve passage R, which is 
equivalent to another nozzle P, except that it enters at the side of the central tube. The 
water is thus sprayed in with the kerosene, a constant stream being sucked up into the 
reservoir from the base, the excess overflowing and going back into the base. The check valve 
S allows the kerosene to come up the tube K, but not to get back. The check valve N allows 
the kerosene to go down the tube Af, but not to get back. In other words, the main piston 
of the engine acts as a pump, valves N and S being equivalent to its suction and discharge 
valves. 

The cage T can be removed from the reservoir by taking out the plug U. It can then be 
lifted out and the valve examined. 



The joint V is held tight by a set screw IT. A cone of wire gauze, A r , is inserted in the 
upper end of the tube, which helps break up the mixture of kerosene, air, and water into a 

Very ^ e B ”^r^e:T“servoir, Y, is placed on the side of the mixer, and has 

a needle valve corresponding to that of the water injection The top of th.s reservoir is covered 
a needle valve cone i 6 simolv lifted out, filled with gasoline, and inverted 

b VtiT r devoirThis“ll mSES the level of gasoline just at the bottom o, 

into the top ot tne . • T t w m be seen that this arrangement provides 

the bottle when it is in place in the es rv^It ^ ^ ^ reservoir % these ■ levels 

a constant level of water ^"discharge openings into the central tube, and therefore when 
being just below the top o • «=> f gaso ii ne water, or kerosene into the mixer. 

T h he and kerosene fed in can be adjusted by their respective needles, 

Z, R, and Q. 




















































































620 


APPENDIX 


The small slide No. 1, on central nozzle P, is raised by pushing up on the rod No. 2, 
when starting the engine. This slide is pushed up into the neck of the mixer at G to throttle 
the flow of air, and cause more vacuum in tube C when turning the engine slowly by hand. 
This draws the fuel in through the gasoline nozzle more readily than when the neck G is wide 
open. In warm weather it is not always necessary to push the slide No. 1 up into place for 
starting. After starting, it must always be pulled down again after the engine has taken an 
explosion or two. 

In the casting 0, Fig. 715, a second roll of gauze is introduced into the tube with the 
cap P. On taking this cap out the gauze will come with it, as the two are attached. The 

gauze cone and the roll of gauze above are both used to break up the kerosene into fine mist, 

and prevent any condensation and drip from the tube. 

Starting. Go back and read over the instructions in regard to setting up, so that you 
know everything is right. 

Be sure that there is kerosene and water in the base. Don’t take anybody’s word for it, 

but look yourself, and measure it to be sure you know just how much you have. While there 

is no way for it to leak out, yet if the engine has been damaged in delivery there might be 



Fig. 71G. 


a crack in the sub-base that would allow it all to leak out shortly after you filled it. The 
engine will not run without fuel. 

Be sure that all the working surfaces are oiled; you cannot get too much oil on a new engine. 
Be sure that the sight feed oiler at the top is turned on and is dropping. See that the grease 
cup is turned down so that the hard oil is forced, when warm, through the feed to the bearing. 

_ 0il the small holes on the governor weight and stud. Throw off the switch at F on Fig. 
707. The cut Fig. 717 shows this switch when it is thrown off, and also how it looks when 
it is thrown on. 

To Start on Gasoline. Revolve the engine to starting position; that is, so the piston is 
beginning the forward, or explosion, stroke. To do this easily, relieve the compression by holding 
the inlet valve R (top valve R, Fig. 713) down during the compression stroke; crank the engine 
just past the compression stroke to the beginning of the explosion stroke (the switch being open); 
let the wheel stand in this position for a minute, w r hile you take the gasoline cup, No. 3, Fig. 





















































































































SOME DETAILS FROM PRACTICE 


621 


716, which is part of the mixer, and fill it with gasoline. When it is filled, turn it upside down 
in the reservoir Y, Fig. 716; open the gasoline needle Z, Fig. 716, half or three-quarters of a 
turn. See that the needle valves for water, R, and kerosene, Q, Fig. 716, are closed. It will 
be well to remember that every moving part must be oiled by some means. Push up the 
rod, No. 2, Fig. 716, as far as it will go. Throw the switch on as per Fig. 717; crank the 
engine over sharply a few times, and it should begin to take explosions, providing the gasoline 
needle valve Z is set right. It may require a little more, or a little less opening of this valve 
to start the engine, and if it does not go the first time, try a little different setting of this 
valve. After the starting position of the valve is learned once, it can be put in this position 
for starting every time. 

As soon as the engine begins to explode, pull down the rod No. 2, Fig. 716, at once. If 
this rod is left up after the engine is started, the mixer will flood; that is, take too much 
gasoline to make an explosive mixture, and shut down. As soon as the machine is running, 
the valve Z, Fig. 716, can be regulated until the engine is taking the sharpest explosions. 
When this needle is set right, the machine will run up to speed quickly, and begin to cut out 
explosions by means of the governor; that is, when it gets up to speed it should take an 
explosion, cut out several, and then take another. If the exhaust from the engine is black, 
too much gasoline is being used, and the needle Z should be closed a little; if a popping sound 
is heard at the engine, too little gasoline is being used. A bluish-white smoke coming from 
the exhaust means too much lubricating oil, but on a new machine this should be left this 
way for a while, and the oil should be cut down only after the engine has been running a 
day or so. It is better to have a little too much lubricating oil coming into the engine to 



Fig. 717. 


start with and only after some considerable running should this be cut down to where it does 
not show a bluish-white smoke. Eight drops per minute is about the usual feed after the 
engine is worked in and running right, and with eight drops per minute the colorless exhaust 
will be found. In starting a new engine, at least fifteen drops per minute should be used. 

The engine is now running on gasoline, and after it has run for a minute or two, the 

needle valve Q (Fig. 716) may be slightly opened. This is the valve which admits kerosene to 
the engine and should not be opened until the head of the engine is too warm to hold the 
hand on It should be always possible to open up the kerosene valve inside of two minutes 
after the' engine is running. As the valve Q is opened the gasoline valve Z should be closed. 
If the engine does not pick up and run promptly on kerosene, it should be run on gasoline 
a little longer. In opening up the kerosene valve Q, it should be opened from about three- 
quarters to° one turn. After the engine picks up and begins to take explosions on kerosene 
with gasoline valve Z closed, the kerosene valve Q should then be gradually closed down until 

the smoke from the exhaust is only the faintest blue. 

The engine is now running on kerosene, and the load can be thrown on. When the load 
is thrown on after about ten minutes’ running, a knocking will be heard in the cylinder, due 

to the heating up of the engine. This knocking is due to the fact that as the engine heats 

un the kerosene in the engine becomes warm enough to pre-igmte because of the compression. 
At'this time the water admission valve R should be opened. This valve should be opened gradually 
and as it is opened the knocking will become less and less, unit the engine is running smoothly If the 
needle valve is opened too wide, it will cause the engine to slow down, lose power, and final y s op, 
and the proper point can be determined by this. In other words, kerosene needle valve 0 mus be 
regulated to reduce the smoke from the exhaust to the minimum, and still have the engine pulling 
strongly' and the water valve R should be as wide open as is possible without cutting down the 
powef of the engine. None of the valves, after they are once set right, will require any adjust- 
















622 


APPENDIX 


ment for change in load, and the engine will run steadily through all such changes up to its 
maximum power. The setting of these needle valves which will give these conditions, can readily 
be found by experiment, and when this is once determined, the valves should be set at this 
point when running the engine thereafter. The only thing which can change these settings will 
be dirt allowed to get into the kerosene or water compartments through carelessness. If dirt 

gets in, the parts should be removed and cleaned. If running water is used instead of a tank, 

it should be turned on before starting the engine. This water should come away from the 
jacket warm to the hand. If it is warm enough to be uncomfortable, a little more water should 
be turned on. 

. To Start on Kerosene. Read over the instructions for starting on gasoline and be sure 

you understand them. 

On the head of the engine (Fig. 714), is a bulb J. This bulb can be heated with a blow¬ 
torch to about red heat, and the engine started on kerosene, without batteries, magneto, or 
gasoline. In this case, of course, the gasoline valve should be closed and the kerosene valve 
opened three-quarters of a turn from the closed position. Torches are made, and can be 

furnished, when required, operating with kerosene so that the only fuel needed for starting or 
running is kerosene. It is much easier, however, to start the machines with gasoline where 
it is obtainable. To start the engine with gasoline requires only a few teaspoonsful before it 
will run on kerosene. With a hot bulb, of course, the engine will start off on kerosene. When 
starting engine on kerosene by means of hot‘bulb, rich mixture must be used to get the first 
explosions. 

Good results will be obtained by opening kerosene needle valve two turns and raising air 
throttle. 

Hot tube must be heated to a dull red heat. 

When using kerosene blast torch the heat of the flame may be increased by raising the 
air pressure. 

To Stop the Engine. Always close water valve R first (Fig. 716); close off the lubricators, 
Fig. 713; close off the kerosene valve Q, Fig. 716. If running water is used for cylinder jacket, 

close off this jacket water; and lastly, throw off the swatch on the 
battery box F, Fig. 707 (the off position is shown in Fig. 717). 
It is quite necessary to close the water and kerosene valves on the 
mixer before throwing off the switch so that any w f ater or kerosene 
remaining in the engine will burn out, and not be left in the engine 
to rust or cause hard starting. 

What to do if the Engine does not Start. Look everything 

over carefully. See that you have followed out the instructions 
exactly. It may be that the gasoline needle valve is too wide open, 
or too far closed to give proper mixture. If this is not found to 
be the case, test the spark plug as per instructions on page 624, 
and if the trouble is not there, test the coil (p. 626). 

If the engine, while running, begins to miss fire and explosions 
occur occasionally in the exhaust pipe, the trouble is usually in the 
ignition system. If the ignition system is right, and the exhaust 
becomes too smoky, too much fuel is being used; if a popping occurs 

in the mixer at the engine, too little fuel is being used. If the 

spark is w r eak at the spark plug, but occurs regularly, the trouble 
may be in the adjustment of the thumb screw on the coil (see 
p. 623), or there may be a loose connection in the battery box, 
or the batteries may have run down. If a magneto is used, this is 
eliminated, for the magneto will always give current, if the brushes are clean. It is possible that 
the kerosene has been used up, and if the engine shuts down the trouble is usually that it 
has run out of fuel. This is the first thing to look for when trouble occurs. 

If you have plenty of kerosene and water, unfasten the wire at E, Fig. 707, and unscrew' 
the spark plug, taking it out of the engine. Then connect the spark plug up again with the 
W'ire and lay the plug on its side, on the metal of the engine, being sure that the point does 
not touch any metal, and that the wire going to the plug does not touch the engine. Throw 
on the switch again and turn the fly-wheel slowly to the right. When you reach the sparking 
point (which may take two turns of the w'heel), you should hear a buzzing noise and there 

should be a stream of sparks two or three times the size of a pin head, between the point and 



Fig. 718. 














































SOME DETAILS FROM PRACTICE 


623 


outside metal of the spark plug. If the plug sparks down the outside or high up on the inside, 
or the spark does not jump between the point and the metal of the plug, the plug is dirty or 
the insulation is cracked. Take your plug apart and clean (instructions, p. 624). If you have 
the switch on, the connecting wires tight, and have followed the above instructions and on 
turning the fly-wheel over to the right fail to hear the buzzing noise from the spark coil, there 
is a loose wire somewhere in the cells, the batteries are exhausted, or the spark coil needs 
adjustment. The fault lies so very seldom with the spark coil that we do not recommend it 
to be touched at all, except as a last resort, as in nine cases out of ten the fault lies else¬ 
where. Do not touch your coil unless you positively know that the trouble is there and nowhere 
else. The chances are* that it is somewhere else. If the trouble was with the spark plug, and 
you now get a good spark, put it back, connect the wires, and your engine will probably start 
right off. 

If you get no spark, and the plug appears to be right, look at the batteries. 

Make sure first that there is no loose connection, that all thumb nuts are tight; a loose 
wire anywhere will be enough to stop the engine. If you have recently changed cells, look 
over the wiring and be sure that it is wired up right, as per Fig. 707. If everything is right, 
take your battery tester, and test each cell for its amperage—each dry cell should show at 



least 14 volts and 15 amperes. It will not do satisfactory work unless it shows at least 8 
amperes If it does not, throw it away and put in a new cell. 4ou are simply destroying 
“ur good cells by leaving this bad one connected up with them. Every user of a gasoline 

engine a „ righti if the wires are right, and the plug is all right, squirt a little 

rSe "l^may 1 MlLJt " “iMt 

engine will take one or two The feed may be stopped up, 

a 

before it was sent out, in the eng it does not start then, let it stand a 

it is flooding because of some than or dinary-if it still does not start, 

few minutes, try with the needle , jP t tlle right- the coil is ordinarily the last 

look at the coil, and turn the d neve be’ touched until it is' positively 

thing to get out of order, and for h s reason^ shoutdm ^ ^ ^ & ^ when 

known that the trouble lies he e. screw G n the side of the frame just below large 

the sparking point is reached, !o° . . d t again . if it then does not start, turn 

screw B, Fig. 719, turn screw B down a little ana tn again. 
















624 


APPENDIX 


down screw B a little further, then try it, always remembering that the vibrator A must still 
have space in which to move. If it refuses to work, your batteries are dead, or you have a 
loose wire, or your spark coil has been spoiled by having a nail driven into it, or being 
injured in transit or unpacking. The chances are ninety-nine out of one hundred that it is a 
loose wire or a dead cell. Be sure that you have tested the batteries and wires before laying 
the difficulty to the coil. 

If magneto is used the spark can be tested by taking the plug out, connecting it to the 
wire again, and laying it on some metal part of the engine. Crank the engine over sharply, 
watching for the spark between the points of the plug. If the spark does not occur, the same 
reasons may hold good as in the case of the batteries, except that in place of the batteries 
being run down, the brushes of the magneto may be dirty. 

How to Clean a Spark Plug. If the spark plug is dirty, wipe it off carefully with a dry 
cloth, and scratch the points with a knife so that they are bright and clean. Remove all 
carbon and oil from the porcelain. The points should be about of an inch apart; if they 
are not, bend them carefully with a pair of pliers. 

If the spark takes place down the outside, or anywhere except directly between the point 
and the side, it is dirty or the insulation is cracked. Take the plug apart. You can do this 


/rs/evc/ms /or /Jd/os/by Go terror L o/th. 


YY/tert roffor /S 0/7 hjybpc //?/ of oc/rr /be jorer/ror 
/o/ch *sboit/</ c/cor fbe fc/e/.yor buffo/; by f z fofc 0/00 



to £ 


r 'J A 


/Pber Ti/rr/ry ery/re orer by bore/ /be yoterror 



easily by using two wrenches. Be sure that the insulation is not broken. If it is, you must 
get another plug. Either put in new porcelain or get a new plug. 

Instructions for Cleaning Mixer and Its Valves. In Fig. 716, to clean the mixer, unscrew 
the screw IF; take out the plug U ; lift the cage T out of the reservoir; and unscrew the cap 
No. 4 from the end of the cage T. The small valve N can be easily cleaned. Put the valve 
and cap No. 4 back into place; and the cage T back into the reservoir, after having cleaned 
out any sediment that may be deposited in the bottom of the reservoir. Be sure that the 
little gasket V that the cage rests on is clean, as this joint must be tight after the plug U 
and set screw IF are in place and screwed down. The valve S can be cleaned by taking off 
the back nut which holds the pipe tight. Unscrew the whole valve from the base, and unscrew 
the cover of the valve from the body of same. The valve can then be thoroughly cleaned on 
the inside, put together and screwed into place again. To clean the needle valves from the 
reservoirs to the mixer, they can be simply unscrewed until they come out. It is well to wash 
out the holes with gasoline or kerosene when the needles are out so that any dirt in the holes 
will not be left in the threads. The small plug No. 5, Fig. 716, can be removed and the hole 
to the needle valve thoroughly cleaned. When this has all been done the cone X should be 
removed and heated in a flame to burn off any lint which may have accumulated on the gauze. 
This gauze should be heated in a clean flame so as not to leave a carbon deposit. The'cover 
P, Fig. 715, can be removed. To this cover is attached a roll of gauze, and this should also 
be heated slightly, enough to burn off the lint, but not to melt out the babbitt in the cap, 
afterwards being put back into place. Where an engine has worked in a dusty place, such as 
running a threshing machine, a very considerable amount of dust is carried in the'air, and 
clogging may occur once in a while. It will usually show by loss of power in the machine; 
when all the other parts of the engine are in good working order. 


SOME DETAILS FROM PRACTICE 


625 


How to Adjust the Governor Latch. If the engine does not govern properly, look at the 
governor latch. When the roller is on the high point of the cam, the governor latch should clear 
the governor button by yg to y6 °f an inch. (Fig. 720.) When turning the engine over by 
hand, the governor weight should lift the governor latch off the button ^ to l of an inch. 
(Fig. 721.) 

How to Adjust the Governor to Regulate Speed. In order to regulate the speed of the 
engine it is necessary to adjust the governor. (Fig. 722.) Screw up the nut M on the governor 
weight to make the engine run faster, and unscrew it to make it run slower. If you wish to 
change the speed of the engine slightly, increasing or decreasing it 50 to 75 r.p.m., this can be 
done while the engine is running by screwing the thumb screw N (Fig. 722) out to increase, 
and in to decrease the speed. Only a very slight turn is necessary. 



Fig. 722. 


How to Tighten the Connecting Rod. If there is a pound or knock in the connecting rod 

it should be tightened either at the piston end or at the crank, or both. To do this you will 

have to remove the oil shield between the wheels by unscrewing two little studs on either 
side, which will allow you to take the shield off. 

In the tool box you will find a special length socket wrench (No. 18) which fits the adjusting 
nut on the top side of the connecting rod at the piston pin inside the piston. Place this socket 
wrench on the nut, take a wench and turn it to the right, turning the socket wrench to the 
right, being careful not to get it too tight. There must be a little play in this connecting rod, 

but not too much. You can tell how much there is by taking hold of the connecting rod and 

shaking it. It must be a trifle loose. Screw this up enough to take up the play which makes 

the noise. It must not knock when running. . . 

On the other end of the connecting rod you will find a split box or bearing which is 
adjusted by four nuts, two top and two bottom. Loosen the lock nuts there and screw up the 
others. If you cannot get this tight enough, take off the cap and cut down the leather shims 
found between the cap and the connecting-rod, then screw' dowm nuts until tightness can be 
felt on turning the crank, then let back a little to loosen. See that the bearing does not heat; 
if it does the box is too tight and should be loosened further. Be sure to keep these two 
adjustments tight enough so that you will hear no knocking. 

















626 


APPENDIX 


Compression. In turning over the wheel when the spark plug is screwed in, you will find 
one point on every second turn w-here the engine can only be turned with difficulty. This 
is caused by the compression which you should always have. If you do not have it, the rings 
on the piston are gummed up and stuck. Give your engine plenty of oil, and the rings will 

probably free themselves, and your compression come back to what it should be. This will 

often be found to be the case where an engine has been standing for some length of time. 

The rings become gummed or rusted into the piston and do not spring out as they should 

and close the space between the cylinder and piston walls. If the compression continues bad, 
disconnect the connecting rod and pull the piston out of the engine, being careful not to injure 
the piston or break the rings. If you have been using good oil your piston will be bright and 
smooth—bad oil, dark and gummy, with the rings stuck up. 

Clean the piston thoroughly with kerosene, be sure that the rings are free before putting 
back, and oil thoroughly with good gas engine oil. In putting the piston back into the cylinder, 
be careful not to pinch or break the rings. If the rings on the piston are loose and the piston 

is bright and smooth and evidently in good shape and you do 
not get compression, it is because the valves are not closing 
properly. It may be that some dirt has gotten underneath 
them, or a little piece of waste, which prevents their closing. 

It is best not to tamper with the setting of the valves 
as they cannot get out of order under ordinary circumstances. 
The accompanying figure (Fig. 723), will show you how they 
should be set, if you ever have occasion to take the engine 
down. 

How to Adjust Spark Coil. If your batteries have grown 
weak and the buzzer fails to work, you can often get a 
spark and go on running for a short time until you can get 
new* batteries, by loosening the set screw on the side of the 

frame below set screw" B, Fig. 719, and then turning down 

screw B and decreasing the space between vibrator A and 
screw r B. Always remember that the longer the space be¬ 
tween the vibrator A and the point of screw B, the 
stronger and hotter will be the spark in the engine, and 
the better the engine efficiency and pow T er. Also, always 

remember that vibrator A must have room to move, and 
that if B is screwed down tightly the engine cannot run at 

all, and that the longer the “ throw ” wffiich you can give the vibrator A the better the engine 

will run. 

On putting in new batteries unscrew screw" B on the coil until the current will just throw 
the vibrator. You can find this point by turning the engine over slowly, everything being 

connected up and the switch on. , 

After months of running the point on screw B may become battered—take it out and 

smooth the point up carefully with a file, so that it gives a good even contact with a clean 

flat point about of an inch in diameter. It is well to keep the top of your coil covered up, 
as the dust may get in and fuse on to the point B and stop your engine. 

General Suggestions. Keep Your Engine Clean. Before starting the engine always oil the 
piston and parts with a hand oil can; see that plenty of standard gas engine oil is in the oil 
cups—do not use any other kind of oil but gas engine oil. 

Always see that all the nuts and screws are kept in their places and are tight. A loose 
nut spells trouble. 

Be sure to oil the rocker arm and valve occasionally. (L, Fig. 713.) If this arm does not 
w r ork freely at first, wash it with a little kerosene, as it may get gummed; then oil it, being 
careful that all kerosene is out. 

Never touch the spark coil until you are sure the trouble is there and have been over 
everything else. It is the last thing to get out of order, and the last thing that should be 
tampered with. 

If the engine turns hard, it is not getting enough lubricating oil, or the oil is poor. 

Put in nc screws or nails that will reach through the outer casing of the battery box and 
the spark coil box; they will ruin the coil. 

Keep the coil in as dry a place as possible; moisture will cause leakage of current. 


















SOME DETAILS FROM PRACTICE 627 

Never leave the switch on. If the switch is left on, it will use up a good set of batteries 
in a few hours. 

In making adjustments never leave the engine in such a position that the buzzer will keep 
working; it will use up your batteries very quickly. In testing or looking for any difficulty, 
if you wish to try the buzzer, do not keep it buzzing any longer than absolutely necessary. 

Keep an extra spark plug on hand, clean, perfect, and ready for use—it will save you time 
and trouble. 

In cleaning up a spark plug, be sure that the insulation is not cracked. 

If a battery becomes exhausted, it will start the engine all right in the morning, but will 

run down before the engine has run long. 

Test your batteries. After the engine has stood for a time the batteries get stronger, but 
they will die on very short running. 

If the engine makes a popping sound at the point of air intake, or air throttle, it is not 
getting enough kerosene. Open the needle valve slowly until this stops. 

If an explosion takes place in the muffler, do not be alarmed; it only means that there 
is too much kerosene, or the spark is poor. 

If the exhaust is bluish-white, do not give the cylinder quite so much lubricating oil or 

kerosene; it may be caused by either. Always remember, however, that of the two faults it 

is better to give an engine too much lubricating oil than too 
little, and especially remember that a new engine needs plenty 
of oil until all the parts are worn perfectly smooth. 

Every user should have a battery tester. This will enable 
you to test your cells yourself, and you will always know 
whether you have a good cell or a poor one. The method 
of applying it is to place one point on one of the screws 
that comes out of the cell, and the other on one of the 
brass tips, press the button on the tester, and note the regis¬ 
tration of the amperes. If it registers over 10 or 12, the cell 
is all right. Test all your cells in the same way. It is not 
necessary to disconnect them from each other to do so. Remem¬ 
ber that if even one cell is weak it reduces the efficiency of 
the whole set. Throw it out and get another. If you cannot 
get another immediately, throw it out and run with the re¬ 
maining cells. 

A battery will wear out in time whether it is used or 
not, so it will not do any harm to have some fresh ones 
on hand. With this battery tester you can also test the cells 
that you buy, and be sure that you get good live ones. Do 
not buy dead ones, as they are worthless to you. 

Keep your engine clean. \ ou cannot get good work from 
a dirty engine. 

Use only good gas engine oil. Steam cylinder oil, or 
sewing machine oil, or any oil that is not made for heat, 

will not do. . 

Always see that all screws and nuts are tight before starting up for a day s run. They 
will work loose on any machine, and keeping them tight will greatly prolong the life of an 

engine. . . 

In Cold Weather. Always drain the engine when through running, and avoid all possibility 
of breakage because of freezing. The engine may be harder to start in cold weathei than in 
warm. In bitter cold weather the metal in the mixing chamber gets so cold that the gasoline 
does not vaporize, but remains liquid and the mixture is not right for the first explosion. If 
the engine does not start, take out the plug and squirt a little gasoline into the plug itself. 
The type “ AK ” engine will always start readily in any kind of weather if it is in adjustment. 
In bitter cold weather when gasoline will not vaporize, the above suggestions may be of value. 

If it is near your house, an equally easy way is to pour some warm watei o\er the inlet 
chamber where the fuel goes into the engine it will then always start readily. 

Remember that in cold weather lubricating oil does not run readily, and take paiticular 
care to see that the sight feed oiler at the top of the engine is letting in the usual quantity 
of oil. 














628 


APPENDIX 


The oil cup should be kept wider open during cold weather to insure proper lubrication. 
Bluish-white smoke from the exhaust will tell you when you have too much. Failure to give 
oil enough may cut your piston badly in a few minutes. 

In order to drain the water out of the engine, open plug at (18), Fig. 712. The water 
jacket of an Olds engine is removable and can easily be replaced if broken by freezing, but to 
let it freeze is simply carelessness. If you drain your engine it cannot happen. Water must 
also be drained out of the base, at point S, Fig. 715. 

Portable Engines. These instructions apply to portable engines, but in addition there is 
a grease cup on the rotary pump which must be kept full of grease and in working order. 
This should be taken care of in the same manner as the grease cup on the engine, saving you 
the bother of greasing in the old-fashioned way. 

In General. Do not be afraid to write to us in case of any difficulty, but above all study 
this instruction book carefully until you have mastered it and understand the engine completely. 

The Olds is the simplest form of engine on the market, and the easiest kind of engine to 

take care of. . 

One of our best repair men goes out on the road with the sole equipment of a lead pencil, 
and we have never yet found him unable to make good. The only thing that can put an engine 
out of business if you fully understand these instructions and follow them—is the actual breaking 
of a part. 

If you are stuck, ask the nearest agent, who can probably set you straight. If he cannot, 
write us; that is what we are here for—to see that Olds engines ruft and run right, and we 
are always glad to know how you are getting along and see that your engine is giving good 

service. .... 

Above all, remember that the engine has run, and will run, and it is simply a matter of 
getting everything right. If everything is right, the engine must go. The repair man with the 
lead pencil is an object lesson; he knows where the trouble is, and it is only a question of 
slight adjustment. Remember that the Olds engine is the simplest gas engine made; it takes 
only a little patience and common sense to work out your difficulties yourself. There are 
thousands of them being used every day in a dozen different countries without any trouble, 
and there is no reason why yours should not work as well as the rest. 

Now, before trying to start your engine, read this book through again. 

OLDS GAS POWER COMPANY, 
Lansing. Mich., U. S. A- 


2. Directions for Starting the Hornsby-Akroyd Oil Engines 
De La Vergne Machine Company 

1. On engines up to 25 H.P., close the valve supplying water to the vaporizer entirely; 
on larger engines, leave the valve slightly open. 

2. Start the lamp for heating the vaporizer. Place it underneath the vaporizer, allowing 

it to remain for a sufficiently long time to heat the vaporizer. 

3. Fill oil cups, and oil the engine thoroughly. 

4. Set the oil pump to give a slightly longer stroke than No. 1 gauge (or full load pump 

stroke). ^ 

5. Open the overflow valve at the governor; work the oil pump by hand until the oil 

is seen through the glass gauge to pass freely from the overflow valve. Then close overflow 
valve. 

6. Pump a little oil into the vaporizer; if the vaporizer is hot enough, on opening the 

test cock on top of cylinder, oil vapor will issue in which case cock must again be closed. It 
is important that the vaporizer he hot enough, before starting the engine. 

7. Give one or two strokes to the oil pump by hand, and turn the fly-wheel over until 

ignition takes place. 

8. Reduce oil pump to No. 3 gauge until load is put on. 

9. Open the valve on the water supply to the vaporizer slowly and carefully, so as to 

prevent rapid contraction of the vaporizer walls. 

10. Take heating lamps off. 

Directions for Cleaning the Vaporizer of the Hornsby-Akroyd Oil Engine. If the engine 

runs on kerosene, carbonization in the vaporizer takes place only if tie engine is overloaded 


SOME DETAILS FROM PRACTICE 


629 


or not handled properly. If the engine runs on fuel oil or crude oil, the vaporizer must be 
cleaned after a running period of between 24 hours and two weeks, depending entirely upon 
the quality of oil used, and conditions under which the engine is run. 

When the vaporizer cap is removed for cleaning, no open flame should be brought into, 
or near to, the open cylinder, and the crank should be in such a position that the exhaust 
valve is open. This precaution will allow any explosive mixture which might remain in the 
cylinder to escape through the exhaust pipe. 


3. Operating Westinghouse Vertical Gas Engines—Instructions to Engineer 

Care of Engine. 1. Maintain crank case oil level at the parting of the rod-bolt nuts on all 
engines up to and including 11 X 12 " three-cylinder, ^ in. above the parting of the nuts on 
13X14" three-cylinder, ^ in. up to the lower brass on 15X23" and 18X22" three-cylinder engines, 
and about 3 in. up the lower brasses on 25X30" three-cylinder engines. 

2 . Use Atlantic Refining Company’s Westinghouse gas engine oil, or oil of same quality, 
in two main oil cups and in crank case. 

3. Use Standard Oil Company’s Renown engine oil, Leonard & Ellis Red Star machine 
oil, or other oil that is free from a gumming tendency, on other parts of engine. If mixing 
valve or ignitor stems have a tendency so gum up, use a half-and-half mixture of this oil and 
kerosene (coal oil) to keep them free. 

4. Governor and ignitor stems must work perfectly free. 

5. Igniter trip stem nuts should be so set that, when the stem has just dropped off the 

cam, the buffer washer holds trip stem -fg in. off cam, and when cam has pushed trip stem 

farthest out, there is 3 W in. to rg in. clearance between igniter arms and trip stem nuts. 

6 . Never run engine without cotter pins in inlet valve stem and igniter stems. 

7. Keep exhaust valves tight. When allowed to leak they become burned. If valves are 
tight the compression will be very noticeable when it is attempted to turn engine forward 
rapidly by hand with exhaust starting levers in vertical position. After grinding exhaust valves, 
always make sure that there is from ^ in. to ^ in. clearance below the lower end of upper 
exhaust valve stem, according to size of engine, when on idle part of cam. When this is 
adjusted correctly, the valves will seat quietly but firmly. 

8 . Keep igniters clean and dry, and do not allow them to leak, or they will heat and be 
damaged. 

9. Remove and examine condition of igniters at least once a week and of exhaust valves 

at least once a*month. This cannot receive too much attention. When any leak appears the 
parts must be immediately ground in to prevent ruining them. A good engineer will examine 

these parts often enough to have them always in good condition, as it is easily done. 

To Start Engine. 1. Make sure of gas at engine by lighting it at pet cock close to throttle. 
It should not burn violently, for that indicates that regulator gives too much pressure. It 
should burn about like an ordinary gas jet. 

2 . Set mixing valve adjustments in best position for starting. This should be carefully 
determined after installation, and marked. The best running mixture should also be detei mined 
and marked. 

3. Exhaust starting levers must be in slanting position. 

4. Start oil cups. 

5. Open gas valve and immediately turn engine forward briskly a number of turns, by 
hand, in the case of small engines, or by dropping the air starting valve stem onto its cam 
and then opening the air cock, in the case of larger engines. As soon as explosions are obtained 
and speed is sufficient, shut off air, throw exhaust starting levers to \eitical position, thus 
giving full compression, and the more powerful explosions will quickly speed the engine up. 
If these levers are thrown too soon, the compression will stop the engine. 

6 . As soon as engine is started, turn on jacket water. This must not be delayed, but it 
should not be done before starting, except in cases where jackets have been drained, as a 

warm engine starts much easier than a cold one. 

7. Move mixing valve adjustment levers out together to running position and put on load. 

8 . Engineer must satisfy himself that everything is normal about engine before going about 

any other work. 


630 


APPENDIX 


Running Engine. 1. Main bearings should not get hotter than crank case. 

2. Engine should receive such attention as a steam engine of same size requires. In the 
case of small engines, a good engineer, by using his intelligence and keeping engines in good 
condition, will have much time to spare for other work, provided the same can be done as a 
sid. e 1SSU6. 

3. Jacket water should leave engine not hotter than 150° F., nor colder than 98° (blood 
heat), to obtain the best results. 

4. Engine should not misfire. Open pet cocks above exhaust valves to determine this 
point. 

To Shut Down Engine. 1. Shut off jacket water. 

2. Shut off gas. . . 

3 . Shut off spark. (Be sure that battery is never left short-circuited.) 

4. Throw out starting levers. 

5. Stop two-cylinder engines with both exhaust valves closed; three-cylinder engines uith 
two valves closed. 

6 . Shut off oil. _ . 

7. Keep engines always warm, if possible, but if there is danger of freezing, drain watei 

from jackets and pipes. 

4. Instructions for the Operation of Pressure Gas Producers Manufactured by the Gas- 

motoren-Fabrik Deutz 1 

Operating the Steam Boiler. The water level in the boiler must be kept at the prescribed 
height, and the pressure maintained at about 4 atm. (60 lbs.). Loading or otherwise interfering 
with the valves is absolutely prohibited. The best fuel to use is gas coke, which burns with 
little tendency to soot or smoke and makes a fire easy to control. The use of other fuels is, 
however, of course quite permissible. 

The exterior of the boiler and especially the steam pipe from the boiler to the injector 
at the producer should be carefully covered with some non-conducting material. 

Cleaning openings (hand holes) through which any mud or scale may be removed are found 
both in the top of the boiler and near the bottom of the (vertical) shell. 

All of the apparatus feeding the boiler should be tested at least once a day to see that 
they operate properly. The feed water should be as soft as possible, not showing much tendency 
to throw down scale. Depending upon the quality of the water, the boiler should be cleaned 
and washed out every one to four months. 

Starting the Producer. To start the producer from the cold, open the valve in the purge 
pipe, the poke-holes in the top, the inside and outside covers of the filling hopper and the ash 
pit doors. (In case of coke producers also the protecting and fire-doors). 

Through one of the fire-doors start a wood fire and throw in from time to time small 
charges of coke or anthracite, but only after the preceding charge has burned through to a 
bright glow. After the glowing fuel column has reached a height of about 6 in., close all of 
the doors, after having luted the edges of them with a little soft loam oi clay to make them 
air tight. Before closing the fire doors of the coke producers do not fail to first put the inner 
protecting door in place. 

Next close the filling hopper and the poke holes and then turn on the steam, the pressuie 
in the boiler having been raised in the mean time. The air blast serves to produce a stronger 
blast and causes a more rapid burning up of the fuel column. More fuel is added from time to 
time, as the glow appears at the top of the charge in the producer, until the column has reached 
the bottom of the fuel magazine, which may be ascertained by introducing a rod through one 
of the poke holes. This is the height at which the level of the fuel should be maintained 
during operation. After the column has finally burned through, an attempt should be made 
to light the gas at the test opening above the purge valve. If the gas burns with a steady 
flame, without lifting away from the opening or going out, for about 5 minutes, it is ready 
for use. Before turning the gas into the cleansing apparatus and the gas holder, make sure 
that all seals (at the scrubber, sawdust purifier, etc.) are properly filled with water, and that 


See illustration, Fig. 407. 



SOME DETAILS FROM PRACTICE 


631 


the water supply is turned on for all parts requiring it (top of producer, scrubber, gas pipe 

leading to seal box, etc.). The supply should be so regulated during operation that the water 

leaves the scrubber as cool as possible, the seal box lukewarm and the top of the producer 
hot. All valves to the gas holder may now be opened, next close the valve in the purge pipe. 
In order to test the quality of the gas during operation a quarter inch burner should be con¬ 
nected to any suitable place in the gas pipe and a flame should be kept constantly burning. 

Stopping the Producer. If the gas make is to be interrupted for a short time only, open 
first the valve in the purge pipe, next shut off the steam. During such brief periods the fire 
in the boiler should only be banked, so that operation can be resumed at a moment’s notice. 

If the producer is to be banked for a»longer period, say over night, shut off the steam, 

close the valve between seal box and scrubber, and open the purge valve. In this way there 

will be a constant small air draft on the producer, sufficient to keep glowing the constants 
of the producer. Fire and ash doors remain closed. The fire under the boiler should be dumped. 
At starting rebuild this fire, and if necessary also clean the grate of the producer. 

Cleaning the Producer Grate. The best time for cleaning the grate is in the morning just 
before starting. Fuel which clinkers badly requires cleaning oftener. 

The grate can he cleaned only when the gas-making is interrupted. The steam must therefore 
be shut off, the purge valve must be open, while the valve between seal box and scrubber is 
closed. Open the ash door, carefully clean the grate of ash and clinker by means of poker 
and hook, and take care that as little as possible of the glowing fuel is drawn down and out 

with the ash. . 

In the case of coke producers a false grate is used to help in the cleaning of the grate. 
For this purpose, after the fire doors and the inner protecting doors are removed, a couple of 
transverse bars are laid in recesses in the fire-door frames and across the doors. Bars are then 
forced through the fuel column, resting at each end on the transverse bars. The grate may 
now be cleaned without the danger of having the entire column come down, since the lattei 
now rests on the false grate. After cleaning, the bars are again drawn out, the transverse 
bars lifted away, and the protecting and fire-doors are replaced. 

In the case of fuel showing a tendency to clinker, it is well to poke down any clinker 

adhering to the sides of the producer, through the openings in the top, before again starting up. 

Dumping the Fire out of the Producer. After an operation lasting several weeks, and especially 
after the quality of the gas has shown a steady falling off for some time, the producer should 

be put out of commission and thoroughly cleaned. . . 

After the fire has been drawn, any ash or clinker adhering to the lining must, be removed 
by poking through the top openings, taking care, however, to injure the brickwork as little 

as Possible. . g go faj wom out that gas can find its way through it and between the sand 

filling and the exterior shell, the only remedy is to entirely replace the lining. 

Regulating the Rate of Gasification. To regulate the amount of gas made, a hut 
valve in the steam line leading to the injector is connected by a light chain with the 
of the gas holder As long as the gas holder is not completely filled, the valve is wide open, 
but'^hfn 8 the bell is near* the top, § its further rise closes the 

until the drop of the bell permits a weight to open the valve again. The .rate of the ga. 
tion may within narrow limits also be controlled by adjusting the valve in the line from 

Btea Ch'arel'n S : the Producer. To charge the generator, fill the hopper and close the outer cover. 
Then bv^ raising the counterweighted lever, open the inner cover, and the fuel will drop into 

^ producer ffterthr a! e the outer cover* again and ignite the gases caught in the hopper 
the producer. Airei uns> w i t any cr as that escapes around 

between the two covers. Do not replace the outer co\er, out let a y g 1 escaping 

the edges of the inner cover burn in the hopper, in order to prevent the noxious gases esca E e 


632 


APPENDIX 


renewed (every 8 to 14 days) only when the valves ahead of and beyond the purifier are 

closed. Open next the ventilating valve on the purifier and then take off the cover. After 

renewing the sawdust, open first the valves in the gas line, and leave the ventilating valve open 
until the gas has displaced all of the air in the purifier. 

Scrubber. The scrubber is filled with pieces of coke large enough to prevent their dropping 
through the grate at the bottom. The water supply at the top of the scrubber should be 

regulated so that there will be a stream at least the size of a lead pencil flowing from the seal 
box. 

The scrubber should be cleaned at intervals of from 9 to 12 months, depending upon the 
length of daily operation. For this purpose the plant must be shut down. Close the valve 

between scrubber and seal box, shut off the water supply and take off both the top and bottom 
cover plates of the scrubber. The latter should then be allowed to stand for several hours in 

order to let the gas escape. Next draw out the coke through the bottom opening. After the 

new filling has been put in, replace the covers, renewing the packing wherever it is defective. 

The coke of the old charge may be used under the boiler. 

All cleaning operations should he undertaken only in the day time, and care should he taken 
to see that there shall be no fire, open flame, light, or smoking in the room during that time. The 
cleaning should always he done by at least two men, and on account of the dangerous, poisonous 
quality of the gas, the greatest care should he exercised, especially to see that there is ample oppor¬ 
tunity to get fresh air. 

Starting up the Plant after Installation or after Cleaning. If the scrubber, gas holder, or 

piping has not yet been in operation, or if they have been open to the air for cleaning purposes, 
the spaces will be filled with air which with the first gas made will form an explosive mixture. If 

therefore a plant is to be newly started, it becomes necessary to completely replace the air 

and after it the air-gas mixture by pure gas. To this end open first the valve in the purge 
pipe, so that the gas first made may escape to the open air. Next close this valve and fill the 
gas holder, then stop the steam blower, and allow the gas holder to discharge through the valve at 
the extreme end of the gas main. Repeat this operation two or three times, depending upon the 
size of the holder and the length of the main. Take care that there shall be no open flame of 
any kind near the discharge opening. After this process, the gas may be regularly used for the 
purposes desired. 

Gas Holder. The counterweight for the gas bell should be of such weight that, when the 
bell sinks, the gas in the mains between the holder and the place of consumption is under a 
pressure of about II in. of water. Since the water in the holder slowly evaporates, provision 

must be made to supply the deficiency from time to time. Once a year the entire filling of 

water should be renewed. 

When the water is being completely drained off, provision must be made to supply air through 
a valve either in the gas main or in the top of the bell. The reason for this is that the drawing 
away of the water may produce so high a vacuum in the bell as to cause the atmospheric 
pressure to crush it in. 

The water collecting in the traps of the gas main should be pumped out once a week. If 
the gas holder is not protected against cold, and if there is danger of freezing, a ‘'convection” 
heating furnace must be installed. In order to obtain a good circulation of the water, the 
level in the holder should stand about 4 in. above the opening of the pipe supplying the warm 
water. This point is important and should be especially looked after. Where no very se\ere 
freezing is to be expected, it is also possible to protect the holder by means of some covering 
and to supply steam from the boiler to keep the temperature of the water sufficiently high. 

Special Notice. Producer gas, on account of its content of carbon monoxide, is very 
poisonous and seriously attacks the organs of respiration. All valves and cocks in the gas 
mains, as well as the filling hopper covers and poke hole stoppers, should therefore be kept as 
clean and gas tight as possible. They should all be tested at least once a month and all leaks 
discovered must be remedied. Since the odor of the gas is nowhere near as strong as that 
of illuminating gas, any leaks occurring are not easily noticeable, and the examination of the 
entire plant for leaks should therefore be carried out with special care. 


SOME DETAILS FROM PRACTICE 


633 


5. Instructions for the Operation of Pressure Gas Producers made by Korting Bros., 

Hannover 1 

1. Steam Boiler. The steam boiler must furnish to the injector a constant supply of dry 
steam. The steam pipes leading to the latter should therefore be well covered. The best fuel 
to use is coke. If other fuels are used care should be had to see that the superheater coil 
placed in the upper part of the producer does not come in contact with the fire, as it may 
be seriously injured thereby. 

The smaller sizes of boiler are fed by hand pump from a feed-water tank, the larger sizes 
are furnished with two Korting Universal injectors. Special directions are furnished for the 
handling of the latter. The boiler should be put into operation before starting the fire in the 

* 2. Gas Producer. To start the producer, open the ash pit doors and the valve in the 
purge pipe, and build a wood fire on the grate. As soon as possible add coal, which is best 
done through the filling hopper, and slowly build up the fuel column as high as the natural 
draft available permits. When the column has reached a height of about 6 to 8 in., close the 
lower doors, start the steam blower and add fuel as rapidly as possible without smothering the 

fir0 

The lower doors should be made gas tight with loam or clay. The use of a little coke 
will make the starting of the fire easier. The coke or coal used in regular operation should be 

drv, because water lowers the quality of the gas made. .. , 

" When the height of the fuel bed has reached say 25 in., no more fuel is added until the 
top layer commences to glow. The interior of the generator may be watched by carefully 

lifting the poke hole stoppers in the top, stopping the blower for a moment while so doing. 

When the fuel bed reaches this stage, the gas made should be of about the proper quality. 
To test it lift one of the stoppers in the top slightly and ignite the gas escaping. If the flame 

burns red_not blue—while the blower is working full blast, it may be considered satisfactoiy 

and mav be led to the scrubber, by closing the purge valve. Beyond the scrubber the gas 
which may be lighted at any convenient test opening, should, when of good quality, burn with 
a long steady flame of blue color with a distinct reddish tinge. If, on the other hand the flame 
shows’ faint blue together with a tendency to go out, the gas is of poor quality and mista 'es 
have been made somewhere in the process. 

During further operations more fuel may be added as long as the upper layer always turns 
a dull red. The proper mean height or thickness of fuel bed should be about 30 in. foi hard 
coal a little less, for coke a little more. To obtain a uniform quality of gas, the level of the 
bed should be maintained at about the same point. The location of the level may be deter¬ 
mined at several places by inserting a rod through the poke holes, the covers of which may 

b ® lflLt‘ f onA f0 in th a daTTlO hours the doors at the bottom of the producer must be 
At least . “' he purge valve and shutting off the blast. This operation must 

first the purge valve, next stop the blower The ash 

may then he removed and the grate freed from clinkers. This work should be carefully done 
ma> then De remu ^ r ...fL linin „ as i on g as possible. At the same time the proper 

in order to piese 1 f importance as far as the satisfactory operation of the 

carrymg out of this c leamug “ hoU r, open the purge valve partly and 

p ant rs ™ed ^ ^ sufficient draft to keep it alive. The blast should be 

slaLd ag^ about 15 minutes before work is resumed, if a good quality of gas is des.red at 

that time. , , • )lt ,i rflW t i ie p re j n the boiler, increase the thickness of the 

After work is stopped S > ‘ h pur°-e valve. Under these conditions the 

fuel bed in the producer to about 3 ft ;- “T^hig In the morning the burned down fuel 
fire in the generator vn l ^P ahve^untd “Loug h t to normal Light in from 30 to 45 
column may, yth the ’ , j in t p e morning just before starting the blast, 

minutes. The best tune to draw ^ 0 n, *e top fhLld be water-cooled to keep 

it fromTurning out t^o ra^dly! andThe watJ should also not be shut off over night. 


1 See illustration, Fig. 406. 




634 


APPENDIX 


From time to time, depending upon the care that the producer receives, the charge should 
be completely drawn to free the lining from adhering clinkers. 

3. Scrubber and Piping. The scrubber should be opened about every eight days to remove 
the mud deposited. Special attention should here be paid to the syphon and the overflow pipe. 
When the water is at its normal level, the syphon should dip into the water about .8 in. The 
tubes of the air pre-heater should from time to time be freed from soot, taking off the cover 
plates and using wire brushes to scrape the tubes. 

If, on account of fouling, the piping offers greater resistance to the flow of gas, the quality 
of the latter will deteriorate. 

4. Sawdust Purifier. W here the best quality of hard coal is used, no sawdust purifier is 
necessary. For tarry fuels or those carrying dust, however, the use of such purifiers is 
recommended. The spaces between the wooden grids or trays should be filled with dry sawdust 
packed as little as possible. About once in two weeks the filling should be renewed. If the 
old charge is thoroughly dried it may be used over again. To prevent the sawdust from falling 
through the grids, a layer of short planing chips may first be put down. After the cover of the 
purifier has been fastened down, a quantity of gas is allowed to escape, until it is certain that 
all the air has been driven out. 

5. Gas Holder. The bell of the holder is employed to regulate the quantity of gas made. 

The gas pressure is easily regulated by adjusting the counterweights. Care should be 

taken not to work with too high a pressure, as this interferes with gasification. In general, 
the pressure should not exceed about 3 in. of water. Periodically, about twice a day, the 
water collectors in the pipe line of the gas holder and those at the low points in the gas mains 
should be pumped out by means of the pump furnished for this purpose. 

6. In General. When a plant is first started up, or after any cleaning operations, a 
sufficient quantity of gas should be allowed to escape to the open air through the ventilating 
cocks or valves placed at several points. The purpose of this is to replace any air in the 
apparatus, because air forms an explosive mixture with the gas. It is also recommended, before 
opening scrubber or purifier for cleaning purposes, to blow air through the system by means 
of the steam blast, the producer being cold, to make sure that the laborers will not breathe 
the gas. 

Any leaks in any of the apparatus or piping should be immediately taken care of, as the 
gas is poisonous. 

. These instructions are not to be considered as imposing responsibility of any kind upon 
Korting Bros. 

6 Instructions for the Operation of Suction-Gas Producers made by the Gasmotoren- 

Fabrik Deutz 1 

1. Method of Operation. General Directions. The engine draws air saturated with water 
a apor through the producer filled with incandescent coal or coke and through the scrubber 
and gas tank. The air and steam passing through the incandescent fuel columns are trans¬ 
formed into the gas which serves to drive the engine. 

The air takes up the necessary water vapor by leading it over the surface of water in the 
vaporizer built around the top of the producer and heated by the gases as they pass out of 
the producer. The water level in the vaporizer is maintained at a certain point by supplying 
a constant stream of water, and allowing the excess to flow through a small overflow pipe into 
the air pipe and from here into the ash-pit where it evaporates and helps to keep the grate 
cool. 

Dui ing operation, on account of the suction of the piston, apparatus and piping are under 
a pressure less than atmospheric, so that if any leak occurs or any valve is opened, air will 
flow into the system but no gas will leak out. On this account the gas will, under such circum¬ 
stances, be diluted,. decreasing the capacity of the engine. If the quantity of air leaking in is 
considerable, there is also the danger that an explosive mixture may be formed. 

It is therefore of prime importance to have all piping, valves and machinery gas tight. 

The best means to test the tightness of the entire system is by means of the blower used 
to start the producer. If air is blown into the system each joint may be quickly tested by 


See illustration, Fig. 412. 



SOME DETAILS FROM PRACTICE 


635 


means of a flame. This examination should be undertaken not only after the plant has first 
been erected, but at stated periods also during regular operation. Air may, however, also find 
its way into the piping by improperly handling the valves. These should therefore be handled 
only as laid down in the following directions, and it is important to perform the operations only 
in the sequence indicated. 

The producer room should always be provided with proper ventilation. The coal used should 
be screened over a sieve having about 6 or 8 meshes to the inch, in order to remove the fine 


dust. 

2. Starting the Producer, after Erection or after it has been Cleaned. Before starting, 
the entire plant should be inspected and the engine made ready for starting. Special care 
should be taken to see that the water seal at the scrubber is filled, that there is water in the 
ash pit, and that the vaporizer is full to the overflow. Close the scrubber valve and the water 
vapor valve and open the blower valve, and the ash and fire-doors. Next pull the removable 
chimney down onto its seat at the top and start a wood fire in the generator as in a common 
stove. After the wood is well lighted add coal, and after this is burned through keep on adding 


it slowdy until the producer is filled. 

In order to bring the charge to incandescence more rapidly the blower is brought into 
operation Just after the fire is well started, or at least when the generator is filled. Ash and 
fire-doors must of course first be closed and the top cover put in place. The blower must be 
kept going until the gas burns at the try cock with a steady flame. Next the grate should 
receive a cleaning, for which purpose the blower is stopped, the blower valve is closed, the 
ash pit and fire-doors are opened. The clinkers are next loosened by poking from below 
through the grate and are then removed through the fire-door without drawing any incandescent 
coal. By poking upwards through the fire-doors any clinker adhering to the walls above the 
doors may also be loosened and taken out. During this period the quality of the gas has of 
course been falling off, for which reason, after closing the lower doors and opening the blow 
valve, the blower is again put on until the gas again proves of good quality. 

When the plant is first started, and also when for some reason there is no gas left from 
the last operation in scrubber or piping, gas must be forced through the system by means of 
the blower before the engine can start. To this end open the purge cock leading to the outer air 
at the engine, open the scrubber valve and close the filling hopper. Gas must then be blown 
through until the air is entirely displaced and the gas burns steadily just ahead of the engine. 

3 & Starting the Engine. After making certain that the gas is of good quality at the test 
cock "the engine may be started. For this purpose stop the blower, open the scrubber valve 
and the water vapor valve, close the charging hopper, and start the water supply to the scrubber 
Then start the engine, closing the purge valve at the same time. 

Between the stopping of the blower and the storting of the engine, no time must be lost. AIL 
arrangements required for starting the engine should therefore be made beforehand 

During operation the water supply to the vaporizer should be such that the overflow is 

sufficient to maintain a pool of water in the ash pit. . , , , . 

4. Charging the Producer. The level of the coal should not be allowed to sink below the 
lower edee of the charging funnel. When charging during operation becomes necessary, it must 
be done immediately after lifting the cover, and the latter must be replaced immediately thereafter. 
The hopper or funnel, either during operation or during stand-by periods, should not be open 

for any length of time without putting the chimney in place. . 

5 Stopping Operation. Stop the engine by following the directions given under the corre¬ 
sponding head in the engine instructions, open the purge valve at the engine and also the 
charging hopper. Next close the scrubber valve and the purge valve. The closing of these 
two vafves must not be forgotten in order to retain a good quality of gas in the system beyond 
the generator to aid in the next starting operation. Stop the water supply to the scrubber 
Ind to the vaporizer. Open the air valve so that the producer can get enough air to keep 

the fire alive and close the air-steam pipe. . , . , 

fi Starting after Short Interruption. Starting the producer, the contents of which are 
U. oian & fnllnws- Onen the fire and ash pit doors, loosen the clinkers 

still incandescent, may be done as follows, upon me me F nobble Next close the 

with a poking bar nad remove them, carrying away as little coal as poss ble. Next close the 

Tors again Ln the blower valve and blow air through the producer until the gas again burns 

l ^ ' the test cock After this proceed as directed under _ and 3. 

properly at the t t * Dependin g upon the kind of coal used, it becomes necessary to clean 

the grate at intervals of several hours. This may be done during operation by opening the ash 


636 


APPENDIX 


pit doors and raking the grate from the center to front and back, thus pushing the clinkers 
in front of the fire-doors. If clinkers also form on the sides of the producer above the fire- 
doors, these can only be removed by poking upward through these doors. But the fire-doors 
must remain open only a very short time and the operation must therefore be performed very 
quickly. 

If the generator is badly clinkered up, the only remedy is to draw the fire altogether and 
to clean thoroughly, the clinkers at the sides being removed by poking through the charging 
hopper. This operation had best be undertaken every Monday morning. It is recommended 
not to clean in this manner just after shutting down on Saturday, because the producer would 
cool down too far, and the lining may suffer under the frequent extreme temperature changes. 

8. Cleaning the Piping. The piping should be examined and cleaned once a month. It 
should be noted that the fouling takes place mostly in the bends and elbows and at the joints 
of the piping to the scrubber. To clean the latter places, the flange unions must be disconnected. 

9. The Scrubber. The scrubber is filled with coke of such size that the pieces can not fall 
through the grate at the bottom. The water is supplied at the top and wets the coke on its 
way down. The supply of water must be so regulated that the gas leaves the scrubber com¬ 
pletely cooled down. 

The scrubber should be cleaned once in from 9 to 12 months, depending upon the hours 
of operation per day. 

To accomplish this the plant must be put out of commission. Close the scrubber valve 
and take off the upper and lower covers of the scrubber. After this let the scrubber stand for 
several hours in order to let the gas escape. Next, draw out the coke by means of an iron rake, 
through the lower door, and put in the new charge. Finally, replace the doors or covers, 
replacing any defective packing. 

All cleaning operations should be carried on only in the day time, and no fire, open flame, or 
smoking should be allowed in the room during that time. The work should be done by at least two 
men and ample supply of fresh air should be provided for. 


It is urgently recommended that, independent of any general directions that may 
be given, there should be posted somewhere in the producer room a brief, easily 
memorized special notice, which may read about as follows: 

7. Instructions for the Prevention of Accidents in the Operation of Suction-Gas Plants 

(Guldner-Motoren Gesellschaft, Aschaffenburg.) 

1. The outer cover of the charging hopper must not be lifted unless the inside conical cover 
is pressed against its seat by the counterweight. The simultaneous opening of both covers is 
dangerous and therefore absolutely prohibited. 

2. While charging coal into the upper hopper, or removing clinker by poking through the 
openings provided, the attendant must hold his face to one side far enough so that any flame 
shooting out can not reach it. Another reason for this rule is that it avoids the breathing of 
any gas that may be present in the charging hopper. 

3. The dismantling or taking apart of any of the apparatus of a producer plant that has 
already been in operation, should be undertaken only when the fire in the producer is completely 
dead, and after the system has been thoroughly aired out. In work of this kind the precaution 
to have no open flame whatever, by means of which the gas could be ignited anywhere near 
the plant, should be strictly observed. 

4. The water seals of the Scrubber must always be kept filled with water. The scrubber 
water must be led away only in closed syphons and mains. 



SOME DETAILS FROM PRACTICE 


637 


II. Specifications, Etc. 

1. General Specifications for the Purchasing of Machinery 

(Adopted by the “ Verein Deutscher Maschinen-Fabriken” at Cologne, 28th Dec., 1889, and 10th Jan., 1891, 

and at Hamburg, 27th May, 1893.) 

1. Prices are free on board works; packing and freight are subject to special agreement. 

2. Payment must be made in legal tender at the offices of the manufacturer (or selling 
agent) and at the following rates: one-third of the purchase price at the time of placing the 
order, one-third after the main parts are delivered f.o.b. works, and one-third three months 
after the machine was first put in service. In any case, however, the last payment must be 
made in six months from the second payment, if delay in starting the machine occurs through 
no fault of the manufacturer or selling agent. Payment by monthly installments is permissible, 
but in such case the average monthly rate should be so fixed as to cancel the obligation in 
about the time outlined above. 

3. The seller guarantees construction and workmanship for a period of . months, agreeing 

to replace without charge during that time all parts that prove defective or useless on account 
of inferior material, poor design, or bad workmanship, or to remedy all defects and troubles 
chargeable to the same causes. 

Natural wear and tear is of course not covered by this guarantee. 

4. Delay in delivery for which the seller is responsible entitles the purchaser to a deduction 
of not to exceed ^ of 1% of the purchase price for every full week that delivery is delayed. 

5. Any other claims for damages than those defined in paragraphs 3 and 4 are not 
recognized. 

6. Strikes abrogate any agreements concerning time of delivery. 

7. Before beginning with erection, the foundations, etc., must be furnished and completely 
settled and the place of erection must be protected against stress of weather. 

8. The manufacturer furnishes one or more skilled erectors, as may be agreed upon. The 
purchaser supplies the erectors free of cost, with the necessary help, hoisting facilities, light, etc., 
and also with minor materials like oil, grease, red lead, cotton waste, etc. The helpers remain 
in the employ of the purchaser. 

9. The charges for each erector, besides traveling fare, shall be as follows: $. for each 

day traveled, and S. per working hour, also $. per day and per man for board and 

lodging. If required, the manufacturers or seller may guarantee that the total costs for one 
or more erectors shall not exceed a certain fixed sum. Ten hours is considered a working day, 
except in mines, etc. (“unter Tage ”), where eight hours are considered equivalent to ten hours. 
Overtime and labor on Sunday are charged for as per special agreement. 

10. Foundation plans are furnished in each case without cost, and, where required, also a 
list of parts. 

] | Anv disagreements regarding the interpretation and fulfilment of the contract shall be 
referred to a board of arbitration, to which each party appoints one member. These two then 
agree upon a third member before commencing negotiations. Should the two representatives of 
the parties concerned fail to reach an agreement, each presents a report to the third member 
of the board, whose decision is final. The costs of the process are divided between the parties 
as agreed upon by the board or decided upon by the third member. 

III. Regulations Concerning the Installation and Use of Internal=Combustion 

Engines 

Under this head the German text points out that the installation of gas producers 
in Germany does not in general require government license unless the plant includes 
a closed boiler used under pressure. In this case permission for the installation of 
the boiler must be obtained and it is subject to inspection under government rules 
and regulation. In some localities, however, the authorities class both pressure and 
suction producers under the general head of “ furnaces,” in which case their installation 
is subject to the general building laws. 







638 


APPENDIX 


The installation of oil engines is in general not subject to government regulations. 
In some cases, however, insurance companies prescribe certain measures in the interest 
of safety. 

As far as the translator’s knowledge goes, there is at present in the United States 
no government supervision in the case of either suction or pressure gas producers. 
Insurance companies concern themselves with oil engines, gasoline, and perhaps crude 
oil, but not kex’osene engines, only in so far as to lay down in the insurance policies 
for buildings certain regulations concerning the handling and storing of the oil. For 
these reasons the government and insurance regulations translated rather fully in the 
following subdivisions may be of little interest to the general reader, but they will 
serve to show what points it was thought necessary to cover to make the operation 
of producer gas plants and oil engines safe to both machinery and attendants. 

It is further pointed out in the text that the restrictions put upon the use of 
the lighter hydro-carbon oils, like gasoline, are too * strict, and that the danger of 
explosions is much exaggerated. That the latter is in general the case may be seen 
from the following extract from a report by Dr. Leymann, Wiesbaden: 


“On account of many small fires due to gasoline, kerosene, etc., the idea that these liquids, 
as well as the similar lighter coal-tar oils, like benzol, are explosive in themselves, has gained 
wide acceptance. From extensive information received it appears, however, that although these 
substances are highly inflammable and easily fired, they are not in themselves explosive. This 
fact may also be easily seen by considering their chemical composition, the main constituents 
being carbon and hydrogen. No one has as yet succeeded in firing these compounds either 
through jar or impact, and where explosions have occurred the presence of mixtures consisting 
of aii with the \ apors from these materials has always been proven. The vapors alone 
cannot be ignited, they may be led through incandescent tubes with safety. Even the air-vapor 
mixtures are explosive only under certain stated conditions and within very narrow limits. The 
more volatile the material, tne greater the danger of the formation of explosive mixtures, benzine 
and gasoline heading the list in this respect. The following mixture ratios of vapor to air are 
explosive: for benzol, upper limit 3 * 5 -, lower limit Trl for gasoline, upper limit Tr, lower limit, 
19 > that is, there must be present in the mixtures per unit volume of gasoline vapor not less 
than 19 and not more than 41 unit volumes of air. All other mixtures of the same vapor 
with aii simph burn without explosion, and only those are explosive in which we have, 
expressing the above ratios in per cent, from 2.5 to 6.7% of benzol, or from 2.3 to 4.8% of 
gasoline \ apor. The probability that an explosive mixture may be formed is consequently quite 
small. Concerning closed vessels completely filled with gasoline, therefore, these can explode only 
as a steam boiler explodes, that is, through excess of pressure. It is possible, however, although 
this has been known to occur only once, for an explosion to take place when the vessel is not 
entirely filled, in which case the vapor may form an explosive mixture with the air contained. 
The actual cause of the explosion in the one case cited could not be ascertained. The accident 
vas quite likely due to the same causes that are at the bottom of explosions in air compressors. 
^ essels holding benzine and similar liquids need therefore be considered dangerous onlv in case 
the liquids are in contact with compressed air. Conflagrations, even when large quantities of 
these volatile fuels were concerned, which is, however, rarely the case, have been accompanied 
by explosions but the flames are in general easily controlled by earth embankments and walls. 
Store-rooms for gasoline may advantageously be enclosed by embankments of that kind and 
eA ery karge tank or barrel should be similarly isolated. Since it is hardly likely to avoid spilling 
some of the liquid, through careless handling or otherwise, provision for instant drainage should 
>e made. Finally all gasoline store-rooms, or rooms in which gasoline is used to any extent, 
should be furnished with means of rapidly filling the room with steam, as this will quickly and 
thoi oughly smother any fire. The steam pipe serving for this purpose must be provided with 
a valve outside of the room or building. 


SOME DETAILS FROM PRACTICE 


639 


1. Regulations for the Installation and Operation of Suction-Gas Power Plants 

(Issued by the Prussian Secretary for Commerce and Industry, June 20, 1904.) 

1 . The apparatus for the production and cleaning of suction gas and the gas engine must be 
installed in well lighted rooms at least 11.5 ft. and in the case of engines over 50 H.P., at 
least 13 ft. high. These rooms must have ample ventilation in such manner that the accumula¬ 
tion of gas cannot occur. The rooms must not be used for other purposes. It is permissible 
to install the entire plant in a single room. 

2 . The use of cellar rooms is only permissible when the floor of such cellars is not to exceed 
6.5 ft. below the level of the ground outside. 

3. A direct communication of rooms used for such purposes with living rooms is not ' 
allowed. It is also compulsory to prevent the access of hot air or vapors to living rooms or 
work shops adjacent to or above the engine or producer room. 

4. The floor space of the plant must be of such size that the individual apparatus, engine 
and other accessories may be easily and safely accessible from all sides. Especial care should be 
had to see that pipes do not hinder freedom of movement or interfere in any way with proper 
attendance to producers and engines. 

5. The charging of the producers must be possible without danger to the attendant (from 
fixed platforms, stairs, or ladders). Care should be taken to see that no gas can escape through 
the charge opening into the room. 

6 . The gases produced in the generator during starting up or during stand-by periods must 
be conducted by a gas-tight pipe of proper size above the eaves of the neighboring buildings. 
The exhaust gases from the engine should be carried through a separate iron pipe to the same 
level and in such a way that the neighborhood is not annoyed by the noise. 

7. Arrangements must be made whereby the escape of gas from the producer into the 
rooms during starting up and stand-by periods is prevented. 

8 . Arrangements must also be made whereby, in case of mis-fires or other troubles, the 
explosive mixture is prevented from flowing back into the gas mains, and explosions of such 
mixtures in the exhaust main are rendered impossible. 

9 . Care should also be taken to reduce to a minimum the undesirable features accompanying 
the cleaning of the producer (drawing of ash, removing clinker, etc.). It may be necessary in 
some cases to collect and lead away the hot vapors and gases forming. 

10. Gas washing and scrubbing apparatus should be furnished with instruments (gauges, 
manometers) to indicate the existing pressure. 

11 . The waste water discharged from the washing apparatus must be so treated that it 
leaves the plant without odor and completely neutralized. Ashes and other refuse also must be 
so taken care of as to avoid all annoyance to the neighborhood. 

12 . The ventilating arrangements must not annoy the attendants through drafts or the 
neighborhood through noise or other causes. 

13. If the radiant heat from the gas producers is likely to cause trouble, the producers 

must be suitably covered. Exhaust pipes inside the rooms or buildings must also be either 

cooled or covered. _ 

14. The foundations for engines and other machines in the plant must be so built as to 

prevent annoying the neighborhood with vibrations. 

15. Provision must be made for suitable artificial lighting. 

16. The provisions of the law for the prevention of accidents must be carried out. 

17. The attendants must be furnished with seats and with washing facilities. 

18. In case the gas plant is combined with a storage-battery installation, the rooms for 

the latter must be separate from the former, and must be properly ventilated. The artificial 

lighting of the battery rooms must be furnished only by incandescent lamps with special safety 

hood or by lights from the outside. . , , 

19. The preceding regulations are not to interfere with any rules or regulations already 

established by local building or other laws. 


640 


APPENDIX 


2. Regulations for the Installation and use of Gasoline Engines for Agricultural Purposes 1 

(Issued by the Society of German Fire Insurance Companies.) 

I. Stationary Engines 

A. Regulation Covering Gasoline Engines, the Vaporizer for which is Installed in a Separate Room 

1. The engine must be installed in a room devoid of artificial heat and separated 
from other rooms by walls without openings other than those necessary to pass shafting, 
etc. There must be no combustible materials either stored or handled in the room, and 
the engine itself must be erected on a fire-proof foundation. In cases where this foundation 
does not extend about 12 in. all around the base of the engine, any wooden floor must be 
covered with sheet iron to at least that distance from the engine base. Above the engine no 
woodwork or combustible should be tolerated within 3 ft. of the engine; at the sides this clear¬ 
ance should be at least 18 in. 

2. The exhaust pipe near the engine must be covered by some fire-proof material to remove 
danger of fire from this source. 

3. The room containing the gasoline vapor producer must not communicate with any other 
room, except through the necessary wall openings for the passage of pipes. The walls must 
be heavy, and both floor and ceiling should be stone or concrete. Good ventilation must be 
provided and the room must not be heated. 

4. In case artificial lighting is necessary for the generator room, the light must either be 
furnished from the outside, or incandescent lamps or safety lamps built on the Davy principle 
must be used. 

5. The charging of the generator must be done only in the day time and from a steel tank 
located in the room itself, which tank should not have a maximum capacity exceeding 200 kg. 

( = about 75 gallons) of gasoline. The liquid must flow only through closed pipes and must be 
moved only by enclosed rotary pumps. 

6. No other gasoline except that permitted under Par. 5 may be kept on the property 
insured. 

B. Regulations for Gasoline Engine Installations in which the Vaporizer is Installed in the Same Room 

1. Engine and vaporizer may be installed only in a separate room containing no 
combustible material. The room must be properly ventilated and must not be heated. The 
room must be separated from adjacent rooms by heavy walls without openings and must have 
fire-proof floor and ceiling. Not even passages for belts or rope drives are permitted in the 
walls separating the engine room from the others, but shafting may be allowed to pass through, 
provided it is closely surrounded at the wall with fire-proof material. 

2. The engine and vapor generator room can be artificially lighted only by light transmitted 
from without, or by electric glow lamps whose closely fitting globes must also cover the lamp 
sockets, or by safety lamps constructed on Davy’s principle. Any other handling of light or 
fire in this room is absolutely prohibited. 

3. To operate the engine only electric ignition must be used. 

4. See Par. 2 under A. 

5. See Par. 5 under A. 

6. See Par. 6 under A. 

C. Regulations Covering Gasoline Engines which do not Employ Vaporizers 

1. See Par. 1 under A. 

2. The gasoline tank belonging to the engine (that is the tank containing the supply for 
the engine and for the ignition flame) must be located outside of the engine room in a separate 


1 These regulations also apply to those engines which use ligroin, naphtha, or any other oil whose 
flash-point is below that of kerosene. 



SOME DETAILS FROM PRACTICE 


641 


room neither artificially lighted or heated, and in which any handling of light or fire is strictly 
prohibited. The supply room must have heavy side walls, fire-proof floor and ceiling, and the 
walls can only be broken through for the passage of the small supply of pipes leading to the 
engine. 

3. See Par. 2 under A. 

4. If alcohol is used for the ignition flame, only the quantity required for daily consumption 
may be kept in the engine room and it must be stored in a tightly closed sheet iron or steel 

tank. 

5. Concerning the filling of the gasoline supply tanks, see Par. 5 under A. 

6. See Par. 6 under A. 


II. Portable Engines 

A. Regulations Covering Gasoline Engines with Separate Vapor Producer 

1. Portable engines (locomobiles) must be placed at least 15 ft. from any buildings, 
barns, sheds, and threshing machines, and no easily combustible material must be allowed 
to get any nearer than this distance. 

2. Only electric ignition must be used in operating the engine. 

3. See Par. 2 under I, A. 

4. The amount of gasoline in storage must not exceed 200 kg. (about 75 gallons) and must 
be kept in a sheet steel tank. The place of storage should be either a substantial room or 
building with fire-proof floors and ceilings, well ventilated but not heated, and having no com¬ 
munication with other rooms j or a pit in the ground covered with an iron plate and located 
at least 100 ft. from any buildings or combustible material. 

5. The engine vaporizer must be filled only in the day time direct from the storage tank 
by means of closed ducts or pipes and enclosed rotary pump. I* or this purpose the engine must 
always be moved to the place of storage. 

The vaporizer must also be emptied in the day time ana only through a screw connection 
with the storage tank mentioned under Par. 4 above. 

6. The engine can only be placed in its shelter on the property after the \ aporizer has 
been completely drained. 

B. Regulations Covering Gasoline Engines which do not Employ I aporizers 

1. See Par. 1 under II, A. 

2. See Par. 2 under II, A. 

3. See Par. 3 under II, A. 

4. See Par. 4 under II, A. . 

5. See Par. 5 under II, A, except that “gasoline tank" should be substituted for ‘ vaporizer. 

6. See Par. 6 under II, A, except that “gasoline tank” should be substituted for “vaporizer.” 


3. Special Regulations Covering the Use of Portable Kerosene Engines in the Agricultural 

ndustries 

(Issued by the Society of German Fire Insurance Companies.) 

1. See Par. 1 under 2, II, A above. 

2 The fuel used must be only common kerosene having a specific gravity of at least .SU. 

3! The kerosene tanks may be filled only in the day time, and no kerosene or alcohol 
supplies should be stored in the immediate vicinity (see under 2, II, -1). 


4. Special Regulations Covering the Use of Alcohol Engines in Agricultural Industries, as 

well as for General Industrial Purposes 

(Issued by the Society of German Fire Insurance Companies.) 

1 The room in which the engine is installed must be separated from other rooms by thick 
walls without openings (except those for the transmission), it must have fire-proof floor and 
ceiling, must not be heated or used for any other purpose, and should not contain any com¬ 
bustible materials. 


642 


APPENDIX 


2. No lamps of any kind, except incandescent lamps, must be brought any nearer than 
7 ft. to the engine or any tank or vessel containing alcohol. 

For general industrial purposes, paragraphs 1 and 2 above show a somewhat different form: 

1. The engine must be installed only in a room in which there are no easily inflammable 
materials either stored or handled, and must be placed on a fireproof foundation. If this 
foundation does not project beyond the engine base at least 12 in. on all sides, any wooden floor 
must be covered to that distance with sheet iron. Above engine and alcohol tanks woodwork 
or other combustible materials must be kept at a distance of at least 3 ft. while at all sides 
the clearance should be at least 18 in. 

2. Both engine and vessels containing alcohol must be distant at least 7 ft. from any heated 
stove or heated pipes, and no illuminating agents, except incandescent lamps, should be brought 
any nearer than that distance. 

3. The exhaust pipe near the engine must be covered by some fire-proof material. 

4. The alcohol supply tank must not be placed directly over the engine, but should be 
securely fastened at a distance of at least 7 ft. to one side. The tank or tanks must be made of 
sheet metal, and must be furnished with tight covers. The alcohol must flow from the tank 
to the engine only in closed ducts or pipes. 

5. The supply tank should be filled only in the day time whenever that is possible, but 
where artificial illumination is necessary only incandescent lights or safety lamps must be used. 

6. The only supply of alcohol allowed in the engine room is that required for a single 
filling of the supply tank. It must be kept in tight sheet metal cans and even then should 
only be brought in just before it is needed. 

7. The supply of alcohol in storage must not exceed 250 kg. ( = about 80 gallons). It must 
be kept in a fire-proof chamber or vault or outside of the buildings insured in a place such 
that these buildings can not be endangered. 


IV. Regulations Concerning the Testing of Gas Engines and Gas Producers 

1. Rules for Conducting Tests of Gas and Oil Engine 

(Am. Soc. of Mech. Engineers, Code of 1901.) 

1. Objects of the Tests. At the outset the specific object of the test should be ascertained, 

whether it be to determine the fulfilment of a contract guarantee, to ascertain the highest 

economy obtainable, to find the working economy and the defects as they exist, to ascertain 
the performance under special conditions, or to determine the effect of changes in the conditions; 
and the test should be arranged accordingly. 

Much depends upon the local conditions as to what preparations should be made for a test, 

and this must be determined largely by the good sense, tact, judgment, and ingenuity of the 

expert undertaking it, keeping in mind the main issue, which is to obtain accurate and reliable 
data. In deciding questions of contract, a clear understanding in regard to the methods of test 
should be agreed upon beforehand with all parties, unless these are distinctly provided for in 
the contract. 

2. General Condition of the Engine. Examine the engine, and make notes of its general 
condition, and any points of design, construction, or operation which bear on the objects in 

iew. Make a special examination of all the valves by inspecting the seats and bearing surfaces, 
and note their condition, and see if the piston rings are gas-tight. 

If the trial is made to determine the highest efficiency, and the examination shows evidence 
of leakage, the valves and piston rings, etc., should be made tight, and all parts of the engine 
put in the best possible working condition before starting on the test. 

3. Dimensions, etc. Take the dimensions of the cylinder, or cylinders, whether already 
known or not; this should be done when they are hot, and in working order. If they are 
slightly worn the a\ erage diameter should be determined. Measure, also, the compression space 
foi clearance volume, which should be done, if practicable, by filling the spaces with water 
previously measured, the proper correction being made for the temperature. 

4. Fuel. Decide upon the gas or oil to be used, and if the trial is to be made for maximum 
efficiency, the fuel should be the best of its class that can readily be obtained, or one that 
shows the highest calorific power. 


SOME DETAILS FROM PRACTICE 


643 


5 CalibratioH of Instruments used in the Tests. All instruments and apparatus should 
be calibrated and their reliability and accuracy verified by comparison with recognized standards. 
^ ppaiatus liable to change or become broken during the tests, such as gauges, indicator springs, 
and thermometers, should be calibrated both before and after the experiments. The accuracy of 
all scales should be verified by standard weights. In the case of gas or water meters, special 
attention should be given to their calibration, both before and after the trial, and at the same 
rate of flow and pressure as exists during the trial. 

(a) Gauges. For pressures above the atmosphere, one of the most convenient, and at the 
same time lcliable, standards is the dead-weight testing apparatus which is manufactured by 
man ) r the prominent gauge makers. It consists of a vertical plunger nicely fitted into a 
cylinder containing oil or glycerine, through the medium of which the pressure is transmitted 
to the gauge. The plunger is surmounted by a circular stand on which weights may be placed, 
and by means of which any desired pressure can be secured. The total weight, in pounds, on 
the plunger at any time, divided by the area of the plunger and of the bushing which receives 
it, in square inches, gives the pressure in pounds per square inch. 

Another standard of comparison for pressures is the mercury column. If this instrument 
is used, assurance must be had that it is properly graduated with reference to the ever varying 
zero point; that the mercury is pure, and that the proper correction is made for any difference 
of temperature that exists, compared with the temperature at which the instrument was 
graduated. 

For pressure below the atmosphere, an air pump or some other means of producing a 
vacuum is required, and reference must be made to a mercury gauge. Such a gauge may be 
a U-tube having a length of 30 in. or so, with both arms properly filled with pure mercury. 

( b) Thermometers. Standard thermometers are those which indicate 212° F. in steam 
escaping from boiling water at the normal barometrical pressure of 29.92 in., the whole stem 
up to the 212° point being surrounded by the steam; and which indicate 32° F. in melting ice, 
the stem being likewise completely immersed to the 32° point; and which are calibrated for points 
between and beyond these two reference points. We recommend, for temperatures between 212 and 
400° F., that the comparison of the thermometer be made with the temperature given in 
Regnault’s steam tables, the method required being to place it in a mercury well surrounded by 
saturated steam under sufficient pressure to give the right temperature. The pressure should be 
accurately determined as pointed out in the above section (a), and the thermometer should be 
immersed to the same extent as it is under its working condition. 

Thermometers in practice are seldom used with the stems fully immersed; consequently, 
when they are compared with the standard, the comparison should be made under like conditions, 
whatever those happen to be. 

If pyrometers of any kind are used, they should be compared with a mercury thermometer 
within its range, and if extreme accuracy is required with an air thermometer, or a standard 
based thereon, at higher points, care being taken that the medium surrounding the pyrometer, 
be it air or liquid, is of the same uniform temperature as that surrounding the standard. 

(c) Indicator Springs. The indicator springs should be calibrated with the indicator in as 
nearly as possible the same condition as to temperature as exists during the trial. This 
temperature can usually be estimated in any particular case. A simple way of heating the 
indicator is to subject it to a steam pressure just before calibration. Compressed air, or com¬ 
pressed carbonic acid gas, are suitable for the actual work of calibration. These gases should 
be used in preference to steam, so as to bring the conditions as near as possible to those 
which obtain when the indicators are in actual use. When compressed carbonic acid gas is used, 
and trouble arises from the clogging of the escape valves with ice, the pipes between the 
valve and the gas tank should be heated. With both air and carbonic acid gas, the pipes 
leading to the indicator should also be heated if it is found that they are below the required 
temperature. The springs may be calibrated for this class of engines under a constant pressure, 
if desired, and the most satisfactory method is to cover the whole range of pressure through 
which the indicator acts; first, by gradually increasing it from the lowest to the highest point, 
and then gradually reducing it from the highest to the lowest point, in the manner which has 
heretofore been widely followed by indicator makers; a mean of the results should be taken. 
The calibration should be made for at least five points, two of these being for the pressures 
corresponding to the maximum and minimum pressures, and three for intermediate points equally 
distant. 

The standard of comparison recommended is the dead weight testing apparatus, a mercury 




644 


APPENDIX 


column, or a steam gauge, which has been proved correct by reference to either of these 

standards. ... . . 

When the scale of the spring determined by calibration is found to vary from the nominal 

scale with substantial uniformity, it is usually sufficiently accurate to take the arithmetical mean 
of the scales found at the different pressures tried. When, however, the scale varies considerably 
at the different points, and absolute accuracy is desired, the method to be pursued is as follows: 
Select a sample diagram and divide it into a number of parts by means of lines parallel to the 
atmospheric line, the number of lines being equal to and corresponding with the number of 
points at which the calibration of the spring is made. Take the mean scale of the spring for 
each division and multiply it by the area of the diagram inclosed between two contiguous lines. 
Add all the products together and divide by the area of the whole diagram; the result will be 
the average scale of the spring to be used. If the sample diagram selected is a fair represen¬ 
tative of the entire set of diagrams taken during the test, this average scale can be applied to 
the whole. If not, a sufficient number of samples of diagrams representing the various conditions 
can be selected, and the average scale determined by a similar method for each, and thereby 
the average for the whole run. 

(d) Gas Meters. A meter used for measuring gas for a gas engine should be calibrated by 
referring its readings to the displacement of a gasometer of known volume, by comparing it 
with a standard gas meter of known error, or by passing air through the meter from a tank 
in which air under pressure is stored. If the latter method is adopted, it is necessary to observe 
the pressure of the air in the tank and its temperature, both at the tank and at the meter, 
and this should be done at uniform intervals during the progress of the calibration. The amount 
of air passing through the meter is computed from the volume of the tank and the observed 
temperatures and pressures. 

The volume of gas thus ascertained should be reduced to the equivalent at a given 
temperature and atmospheric pressure, corrected for the effect of moisture in the gas, which is 
ordinarily at the saturation point or nearly so. We recommend that a standard be adopted for 
gas-engine work, the same as that used in photometry, namely, the equivalent volume of the 
gas when saturated with moisture at the normal atmospheric pressure at a temperature of 60° 
F. In order to reduce the reading of the volume containing moist gas at any other temperature 
to this standard, multiply by the factor 


459.4 + 60 b 

45974'+T X ~ 


• (29.92 -s) 
~ 29.4 


CALIBRATION OF A WATER METER 
SKETCH SHOWING METER CONNECTIONS ETC. 


in which b is the height of the barometer in inches at 32° F., t the temperature of the gas 

at the meter in degrees F., and s the vacuum 
in inches of mercury corresponding to the tem¬ 
perature of t obtained from steam tables. 

(e) Water Meters. A good method of calibrating 
a water meter is the following, reference being 
made to Fig. 725. 

Two tees A and B are placed in the feed pipe, 
and between them two valves C and D. The meter 
is connected between the outlets of the tees A and 
B. The valves E and F are placed one on each 
side of the meter. When the meter is running, 
the vavles E and F are opened, and the valves 
C and D are closed. Should an accident happen 

to the meter during the test, the valves E and F 

may be closed, and the valves C and D opened, so as 
to allow the feed water to flow directly to the 
point of use. A small bleeder G is opened when the 
valves C and D are closed, in order to make sure that there is no leakage. A gauge is attached 

at H. When the meter is tested, the valves C, D, and F are closed, and the valves E and 1 

are opened. The water flows from the valve I to a tank placed on weighing scales. In testing 
the meter the rate of flow should be the same as that on the test, and the water leaving the 
meter is throttled at the valve I until the pressure shown by the gauge H is the same as that 
indicated when the meter is running under the normal conditions. The piping leading from the 






































SOME DETAILS FROM PRACTICE 


645 


valve I to the tank is arranged with a swinging joint, consisting merely of a loosely fitting 
elbow, so that it can be swung readily into the tank or away from it. After the desired 
pressure and rate of flow have been secured, the end of the pipe is swung into the tank the 
instant that the pointer of the meter is opposite some graduation mark on the dial, and the 
water continues to empty into the tank. The tests should be made by starting and stopping 
at the same graduation mark on the meter dial, and continued until at least 10 or 20 cu.ft. 
are discharged for one test. The water collected in the tank is then weighed. 

The water passing the meter should always be under pressure in order that any air in 
the meter may be discharged through the vents provided for this purpose. Care should be 
taken that there is no air contained in the water. The meter should be tested both before and 
after the engine trial, and several tests be made of the meter in each case in order to obtain 
confirmative results. It is well to make preliminary tests to determine whether the meter works 
satisfactorily before connecting it up for an engine trial. The results should agree with each 
other for two widely different rates of flow. 

6. Duration of Test. The duration of a test should depend upon its character and the 
objects in view, and in any case the test should be continued until the consecutive readings of 
the rates at which oil or gas is consumed, taken at say half-hourly intervals, become uniform 
and thus verify each other. If the object is to determine the working economy, and the period 
of time during which the engine is usually in motion is some part of twenty-four hours, the 
duration of the test should be fixed for this number of hours. If the engine is one using coal 
for generating gas, the test should cover a long enough period to determine with accuracy the 
coal used in the gas producer; such a test should be of at least twenty-four hours’ duration, 
and in most cases it should extend over several days. 

7. Starting and Stopping a Test. In a test for determining the maximum economy of an 
engine, it should first be run a sufficient time to bring all the conditions to a normal and con¬ 
stant state. Then the regular observations of the test should begin, and continue for the 
allotted time. 

If a test is made to determine the performance under working conditions, the test should 
begin as soon as the regular preparations have been made for starting the engine in practical 
work, and the measurements should then commence and be continued until the close of the 
period covered by the day’s work. 

8. Measurement of Fuel. If the fuel used is coal furnished to a gas producer, the same 

methods apply for determining the consumption as are used in steam boiler tests. 

If the fuel used be gas, the only practical method of measurement is the use of a meter 
through which the gas is passed. Gas bags should be placed between the meter and the engine 
to diminish the variation of pressure, and these should be of a size piopoitionate to the quantity 
used. When a meter is employed to measure the air used by an engine, a receiver with a 
flexible diaphragm should be placed between the engine and the meter. The temperature and 
pressure of the gas should be measured, as also the barometric pressure and temperature of the 
atmosphere, and the quantity of gas should be determined by reference to the calibration of 
the meter, taking into account the temperature and pressure of the gas. [See Section 5 (d)]. 

If the fuel is oil, this can be drawn from a tank which is filled to the original lc^el at the 

end of the test, the amount of oil required for so doing being weighed; or, for a small engine, 
the oil may be drawn from a calibrated vessel such as a vertical pipe.. 

In an engine using an igniting flame the gas or oil required for it should be included in 

that of the main supply, but the amount so used should be stated separately, if possible. 

9. Measurement of Heat-Units Consumed by the Engine. The number of heat units used 

is found by multiplying the number of pounds of coal or oil or the cubic feet of gas consumed, 

bv the total heat of combustion of the fuel as determined by a calorimeter test. In determining 

the total heat of combustion no deduction is made for the latent heat of the water vapor in 
the products of combustion. There is. a difference of opinion on the propriety of using this 
higher heating value, and for purposes of comparison care must be taken to note vhetier ms 
or the lower value has been used. The calorimeter recommended for determining the heat of 
combustion is the Mahler, for solid fuels or oil, or the Junker for gases, or some form of 
calorimeter known to be equally reliable. (See Poole on “The Calorific Power of Fuels ) 

It is sometimes desirable, also, to have a complete chemical analysis of the oil or gas. The 
total heat of combustion may be computed, if desired, from the results of the analysis, and 

should agree well with the calorimeter values. . . , t 

In using the gas calorimeter, which involves the determination of the volume instead of the 


646 


APPENDIX 


weight of the gas, it is important that this should be reduced to the same temperature as that 
corresponding to the conditions of the engine trial. The formula to be used for making the 
reduction is that already given in Section 5 ( d ). 

For the purpose of making the calorimeter test, if the fuel used is coal for generating gas 
in a producer, or oil, samples should be taken at the time of the engine trial, and carefully 
preserved for subsequent determination. If gas is used, it is better to have a gas calorimeter 
on the spot, samples taken, and the calorimeter test made while the trial is going on. 

10. Measurement of Jacket Water to Cylinder or Cylinders. The jacket water may be 
measured by passing it through a water meter or allowing it to flow from a measuring tank 
before entering the jacket, or by collecting it in tanks on its discharge. 

11. Indicated Horse-Power. The directions given for determining the indicated horse-power 
for steam engines apply in all respects to internal combustion engines. 

The indicated horse-power should be determined from the average mean effective pressure 
of diagrams taken at intervals of twenty minutes, and at more frequent intervals if the nature 
of the test makes this necessary. With variable loads, such as those of engines driving generators 
for electric railroad work, and of rubber-grinding and rolling-mill engines, the diagrams cannot be 
taken too often. In cases like the latter, one method of obtaining suitable averages is to take 
a series of diagrams on the same blank card without unhooking the driving cord, and apply the 
pencil at successive intervals of ten seconds until two minutes’ time or more has elapsed, thereby 
obtaining a dozen or more indications in the time covered. This tends to insure the deter¬ 
mination of a fair average for that period. In taking diagrams for variable loads, as indeed for 
any load, the pencil should be applied long enough to cover several successive revolutions, so 
that the variations produced by the action of the governor may be properly recorded. To 
determine whether the governor is subject to what is called “racing” or “hunting,” a variation 
diagram should be obtained; that is, one in which the pencil is applied a sufficient time to 
cover a complete cycle of variations. When the governor is found to be working in this manner, 
the defect should be remedied before proceeding with the test. 

Note. —When the engine is governed by the hit-and-miss principle the diagrams taken on one card 
should in any case cover the series of consecutive explosions, and the mean diagram should be used as 
the basis of calculations. 

The most satisfactory driving rig for indicating seems to be some form of well-made para¬ 
graph, with driving cord of fine annealed wire leading to the indicator. The reducing motion, 
whatever it may be, and the connections to the indicator, should be so perfect as to produce 
diagrams of equal lengths, and produce a proportionate reduction of the motion of the piston 
at every point of the stroke, as proved by test. 

To test the accuracy of the reducing motion without making special preparations for a 
thorough examination, it is sufficient to make a comparison between the actual proportion of 
the stroke covered and the apparent proportion measured on the indicator, and see how they 
agree. This may be done on a large engine by making the comparison wherever it happens to 
stop, and repeating the comparison when it has stopped with the piston at some other point 
of the stroke. With an engine which can be turned over by hand, or where auxiliary power is 
provided for moving it, the comparison may be made at a number of equidistant points in the 
stroke. To make the test properly, a diagram should be taken just before stopping, and this 
will serve as a reference for the measurements taken after stopping. The actual proportion of 
stroke covered is determined by measuring the distance which the piston has moved and 
comparing it with the whole length of the stroke, making sure that the slack has all been taken 
up. To obtain the apparent indication from the diagram, the indicator pencil is moved up and 
down with the finger so as to make a vertical mark on the diagram, and the distance of this 
mark from the beginning of the diagram compared to the whole length of the diagram is the 
proportion desired. 

It is necessary, of course, to go through these operations without changing in any way the 
adjustment of the driving cord of the indicator, or any part of the mechanism that would alter 
the movements of the indicator. 

In the manipulation of the indicator it is important to keep the instrument in clean 
condition and preserve it in mechanically good order. Ordinary cylinder oil is the best material 
to use for lubricating the indicator piston for pressures above the atmosphere. It is better to 
have the piston fit the cylinder rather loosely—so as to get absolute freedom of motion—than to 
ha\ e a mechanically accurate fit. In tne latter case, extreme care and frequent cleanings are 


SOME DETAILS FROM PRACTICE 


647 


required to obtain good diagrams. No diagram should be accepted in which there is any 
appearance of want of freedom in the movement of the mechanism. A ragged or serrated line 
in the region of the expansion or compression lines is a sure indication that the piston or some 
part of the mechanism sticks; and when this state of things is revealed the indicator should not 
be trusted, but the cause be ascertained and a suitable remedy applied. An indicator which is 
free when subjected to a steady pressure, as it is under a test of the springs for calibration, 
should be able to produce the same horizontal line, or substantially the same, after pushing the 
pencil down with the finger, as that traced after pushing the pencil up and subsequently tapping 
it lightly. When the pencil is moved by the finger, first up and then down, the piston being 
subjected to the pressure, the movement should appear smooth to the sense of feeling. 

The pipe connections for indicating gas and oil engines should be removed as far as possible 
from the ports and ignition devices, and made preferably in the cylinder head. The pipes should 
be as short and direct as possible. Avoid the use of long pipes, otherwise explosions of the 
gas in these connections may occur. 

Ordinary indicators suitable for indicating steam engines are much too lightly constructed 
for gas and oil engines. The pencil mechanism, especially the pencil arm, needs to be very 
strong to prevent injury by the sudden impact at the instant of the explosion; a special gas- 
engine indicator is required for satisfactory work, with a small piston and a small spring. 

12. Brake Horse-Power. The determination of the brake horse-power, which is very 
desirable, is the same for internal-combustion as for steam engines. 

This term applies to the power delivered from the fly-wheel shaft to the engine. It is the 
power absorbed by a friction brake applied to the rim of the wheel, or to the shaft. A form of 
brake is preferred that is self-adjusting to a certain extent, so that it will, of itseif, tend to 
maintain a constant resistance at the rim of the wheel. One of the simplest brakes for 
comparatively small engines, wdfich may be made to embody this principle, consists of a cotton 
or hemp rope, or a number of ropes, encircling the wheel, arranged with weighing scales or 
other means for showing the strain. An ordinary band brake may also be constructed so as 
to embody the principle. The wheel should be provided with interior flanges for holding water 
used for keeping the rim cool. 

A self-adjusting rope brake is illustrated in Fig. 726, where it will be seen that, if the 
friction at the rim of the wheel increases, it will lift the weight A, which action will diminish 



Fig. 726. 


Fig. 727. 



the tension in the end B of the rope, and thus prevent a further increase in the friction The 
same device can be used for a band brake of the ordinary construction. V here space below 
the wheel is limited, a cross bar, C, supported by a chain tackle exactly at its center point, 
may be used as shown in Fig. 726, thereby causing the action of the weight on the brake to 
be upward. A safety stop should be used with either form, to prevent the weights being 

accidentally raised more than a certain amount. 

The water-friction brake is especially adapted for high speeds and has the advantage of 
being self-cooling. The Alden brake is also self-cooling and is capable of fine adjustment. 

A water-friction brake is shown in Fig. 727. It consists of two circular disks, A and B, 
attached to the shaft C, and revolving in a case, E, between fixed planes. The space between 












































648 


APPENDIX 


the disks and the planes is supplied with running water, which enters at D and escapes at the 
cocks F, G, and H. The friction of the water against the surfaces constitutes a resistance 
which absorbs the desired power, and the heat generated within is carried away by the water 
itself. The water is thrown outward by centrifugal action and fills the outer portion of the 
case. The greater the depth of the ring of water, the greater amount of power absorbed. By 
suitably adjusting the amount of water entering and leaving any desired power can be obtained. 
Water-friction brakes have been used successfully at speeds of over 20,000 r.p.m. 

13. Speed. There are several reliable methods of ascertaining the speed, or the number of 
revolutions of the engine crank-shaft per minute. The simplest is the familiar method of 
counting a number of turns for a period of one minute with the eye fixed on the second hand 
of a time piece. Another is the use of a counter held for a minute or a number of minutes 
against the end of the main shaft. Another is the use of a reliable tachometer held likewise 
against the end of the shaft. The most reliable method, and the one we recommend, is the 
use of a continuous recording engine register or counter, taking the total reading each time 
that the general test data are recorded, and computing the revolutions per minute corresponding 
to the difference in the readings of the instrument. When the speed is above 250 r.p.m., it is 
almost impossible to make a satisfactory counting of the revolutions without the use of some 
kind of mechanical counter. 

The determination of variation of speed during a single revolution, or the effect of the 
fluctuation due to sudden changes of the load, is also desirable-, especially in engines driving 
electric generators used for lighting purposes. There is at present no recognized standard method 
of making such determinations, and if such are desired, the method employed may be devised 
by the person making the test and described in detail in the report. 

One method suggested for determining the instantaneous variation of speed which accom¬ 
panies a change of load, is as follows: A screen containing a narrow slot is placed on the end 
of a bar and vibrated by means of electricity. A corresponding slot in a stationary screen is 
placed parallel and nearly touching the vibrating screen, and the two screens are placed a short 
distance from the fly-wheel of the engine in such a position that the observer can look through 
the two slots in the direction of the spokes of the wheel. The vibrations are adjusted so as to 
conform to the frequency with which the spokes of the wheel pass the slots. 

When this is done the observer viewing the wheel through the slots sees what appears to 
be a stationary fly-wheel. When a change in the velocity of the fly-wheel occurs, the wheel 
appears to revolve either backward or forward according to the direction of the change. By 
careful observations of the amount of this motion, the change of angular velocity during any 
given time is revealed. 

Experiments that have been made with a device of this kind show that the instantaneous gain 
of velocity, upon suddenly removing all the load from an engine, amounted to from one-sixth 
to one-quarter of a revolution of the wheel. 

In an engine which is governed by varying the number of explosions or working cycles, a 
record should be kept of the number of explosions per minute; or if the engine is running at 
nearly the maximum load, by counting the number of times the governor causes a miss in the 
explosions. 

One way of mechanically recording the explosions is to attach to the exhaust pipe a cylinder 
and piston arranged so that the pressure caused by the exhaust gases operates against a light 
spring and moves a register, which is provided for automatically counting the number. 

Note. —An instrument for this purpose has been devised by R. Mathot. The following description 
is from his book on “Modern Gas Engines and Producer Gas Plants”: 

The instrument, Fig. 728, is somewhat similar in form to the ordinary indicator. Its record, however, 
is made on a paper tape which is continuously unwound. The cylinder c is provided with a piston p, 
about the stem of which a spring s is coiled. A clock train contained in the chamber b unwinds the 
strip of paper from the roll p' and draws it over the drum p", where the pencil t leaves the mark. The 
tape is then rewound on the spindle p'". A small stylus or pencil / traces the atmospheric line on the 
paper as it passes over the drum p". In order to obviate the binding of the piston p when subjected to 
high temperature of the explosions, the cylinder c is provided with a casing e in which water is circulated 
by means of a small rubber tube which fits over the nipple e'. This recorder analyzes with absolute 
precision the vork of all engines, whatever may be their speed. It gives a continuous graphic record 
from which the number of explosions, together with the initial pressure of each, can be determined, and 
the order of their succession. Consequently the regularity or irregularity of the variations can be observed 


SOME DETAILS FROM PRACTICE 


649' 


and traced to the secondary influences producing them, such as the action of the inlet and outlet valves 
and the sensitiveness of the governor. It renders it possible to estimate the resistance to suction and 
the back pressure due to expelling the burnt gases, the chief causes 
of loss in efficiency in high-speed engines. Furthermore, the in¬ 
fluence of compression is markedly shown from the diagram ob¬ 
tained. 

The recorder is mounted on the engine; its piston is driven back 
by each of the explosions to a height corresponding with their force; 
and the stylus or pencil controlled by the lever t records them side 
by side on the moving strip of paper. The speed with which this 
strip is unwound conforms with the number of revolutions of the 
engine to be tested, so that the records of the explosions are placed 
side by side clearly and legibly. 

Their succession indicates not only the number of explosions and 
of revolutions which occur in a given time, but also their regularity, 
the number of mis-fires. The pressure of the explosions is measured 
by a scale connected with the recorder-spring. By employing a 
very weak spring which flexes at the bottom simply by the effect 
of the compression in the engine cylinder, it is possible to ascertain 
the amount of the resistance to suction and to the exhaust. It is 
simply sufficient to compare the explosion record with the atmospheric 
line, traced by the stylus /. By means of this apparatus, and of 
the records which it furnishes, it is possible analytically to regulate 
the work of an engine, to ascertain the proportion of air, gas, or 
hydrocarbon which produces the most powerful explosion, to regu¬ 
late the compression, the speed, the time of ignition, the temperature, 
and the like. 



14. Recording the Data. The time of taking weights and every observation should be 
recorded, and note made of every event, however unimportant it may seem to be. The pressures, 
temperatures, meter readings, speeds, and other measurements should be observed every 20 or 
30 minutes -when the conditions are practically uniform, and at more frequent intervals if they 
are variable. Observations of the gas or oil measurements should be taken with special care at 
the expiration of each hour, so as to divide the test into hourly periods, and reveal the 
uniformity, or other-wise, of the conditions and results as the test goes forward. 

All data and observations should be kept on suitable prepared blank sheets or in note books. 

15. Uniformity of Conditions. When the object of the test is to determine the maximum 
economy, all the conditions relating to the operation of the engine should be maintained as 
constant as possible during the trial. 

16. Indicator Diagrams and their Analysis. Sample Diagrams'. Sample diagrams nearest 
to the mean should be selected from those taken during the trial and appended to the tables 
of the results. If there are separate compressions or feed cylinders, the indicator diagrams 
from these should be taken and the power deducted from that of the main cylindei. 

17. Standards of Economy and Efficiency. The hourly consumption of heat, determined 
as pointed out in Article 9, divided by the indicated or the brake horse-power, is the standard 
expression of engine economy recommended. 

In making comparisons between the standard for internal-combustion engines and that for 
steam, it must be borne in mind, that the former relates to energy concerned in the generation 
of the force employed, whereas in the steam engine it does not relate to the entire energy 
expended during the process of combustion in the steam boiler. The steam engine standard 
does not cover the losses due to combustion, while the internal-combustion engine standard, 
in cases where a crude fuel such as oil is burned in the cylinder, does cover these losses. To 
make a direct comparison between the two classes of engines considered as complete plants for 
the production of power, the losses in generating the working agent must be taken into account 
in both cases and the comparison must be on the basis of the fuel used; and not only this, 
but on the basis of the same or equivalent fuel used in each case. In such a comparison, where 
producer gas is used, and the producer is included in the plant, the fuel consumption, which 
will be the weight of coal in both cases, may be directly compared. . . 

The thermal efficiency ratio per indicated horse-power or per brake horse-power for internal- 


























650 


APPENDIX 


combustion engines is obtained in the same manner as for steam engines, and is expresesd by 
the fraction 


_2545_ 

B.T.U. per H.P. per hour’ 


18. Heat Balance. For purposes of scientific re'search, a heat balance should be drawn 
which shows the manner in which the total heat of combustion is expended in the various 
processes concerned in the working of the engine. It may be divided into three parts: first, the 
heat which is converted into the indicated or brake work; second, the heat rejected in the 
cooling water of the jackets; and third, the heat rejected in the exhaust gases, together with 
that lost through incomplete combustion and radiation. 

To determine the first item, the number of foot-pounds of work performed by, say, one 
pound or one cubic foot of the fuel is determined; and this quantity divided by 778, which is 
the mechanical equivalent of one British thermal unit, -gives the number of heat units desired. 
The second item is determined by measuring the amount of cooling water passed through the 
jackets, equivalent to 1 lb. or 1 cu.ft. of fuel consumed, and calculating the amount of heat 
rejected, by multiplying this quantity by the difference in the sensible heat of the water leaving 
the jacket and that entering. The third item is obtained by the method of differences; that is, 
by subtracting the sum of the first two items from the total heat supplied. The third item can 
be subdivided by computing the heat rejected in the exhaust gases as a separate quantity. The 
data for this computation are found by analyzing the fuel and the exhaust gases, or by 
measuring the quantity of air admitted to the cylinder in addition to that of the gas or oil. 

19. Report of Test. The data and results of a test should be reported in the manner 
outlined in one of the following tables, the first of which gives a complete summary when all 
the data are determined, and the second is a shorter form of report in which some of the 
minor items are omitted. 

20. Temperatures Computed at Various Points of the Indicator Diagram. The compu¬ 
tation of temperatures corresponding to various points in the indicator is, at best, approximate. 
It is possible only where the temperature of one point is known or assumed, or where the 
amount of air entering the cylinder along with the charges of gas or oil, and the temperature 
of the exhaust gases, is determined. 

If the amount of air is determined for a gas engine, together with the necessary tempera¬ 
tures, so that the volume and the temperature of the air entering the cylinder per stroke, and 
that of the gas are known, we may, by combining this with the other data, compute the 
temperature for a point in the compression curve. In this computation we must allow for the 
volume of the exhaust gases remaining in the cylinder at the end of the stroke. The tempera¬ 
ture at the point in the compression curve where it meets or crosses the atmospheric line will 
be given by the formula: 


T = 


491.4 V' 

V" + V"' + V"" 


-459.4; 


(A) * 


where T ' is the total volume corresponding to the point where the compression curve meets 
or crosses the atmospheric line; V" the volume of the air at atmospheric pressure entering the 
cylinder during each working cycle, reduced to the equivalent volume at 32° F.; V” the 
volume of the gas consumed per cycle reduced to the equivalent at atmospheric pressure and 
32° F.; and V"" the volume of the exhaust gases retained in the cylinder reduced to the 
same basis. To reduce the actual volumes to those at 32° F., multiply by the ratios of 
491.4-r- (T' + 459.4), where T' is the observed temperature of the air and of the gas used as fuel. 
For the exhaust gases retained in the cylinder at the end of the stroke T' may be taken as 
the temperature of the exhaust gases leaving the engine, provided the engine is not of the 
“scavenging” type. 

Having determined the temperature of a point in the compression curve, the temperature 
of any point in the diagram may be found by the equation 

7\ = (T +459.4)^^-459.4 


. • ( B ) 







SOME DETAILS FROM PRACTICE 


651 


Here T 1 is the desired temperature of any point in the diagram where the absolute pressure 
is Pj and the total voulme V u and P and V are the corresponding quantities for the point in 
the compression line having the temperature T computed from the formula ( A ). 

Formula (P) holds only where the weight of the gases contained in the cylinder is constant. 
It is also assumed in this formula that the density of the gas compared to air at the same 
temperature and pressure is the same before and after the explosion. 

A second method may be employed, provided the air which enters the cylinder is measured. 
This will allow for any difference in the density of the gas before and after explosion, and more 
exact values for temperatures on the expansion curve may be obtained than by the first method. 

In this method the density of the exhaust gases compared to air at the same temperature 
and pressure is computed, assuming perfect combustion, and including the effect of the water 
vapor present; and from this density the volume of the gases exhausted per cycle is determined. 
If this volume exhausted per cycle, added to the volume of the gas retained in the clearance 
space at the end of the stroke, be called V in equation (B), and T be the observed temperature 
of the exhaust gases, this equation may be used for determining the temperature of any point 
in the diagram in the way already described. This method is more complicated than the first, 
as it involves the determination of the theoretical density after explosion, but it possesses the 
advantage that it may be applied to an oil as well as to a gas engine. 

A third method of computing the temperature of the various points in the diagram may be 
employed where analyses of the exhaust gases as well as of the fuel have to be made. This 
method is more complicated than the first, but, in common with the second, it possesses the 
advantage that it may be applied to an oil as well as to a gas engine. 

In applying the third method the volume of the exhaust gases discharged per working cycle 
would be given by the formula: 


V 2 =^(Rw+w), 


(O 


where D is the density of the exhaust gases at their observed temperature, computed from the 
analysis, assuming the vapor of water produced through burning the hydrogen in the fuel to 
be in a gaseous state; R the weight of the air which enters the cylinder per pound of fuel 
consumed per working cycle; the value of R, providing there are no unconsumed hydrocarbons, 
may be computed by employing the formula: 


NC 

.33(C0 2 + C0) ’ 


(D) 


where N, CO, and CO represent the proportions, by volume, of the several constituents of the 
exhaust gases, and C the weight of carbon consumed and converted to C0 2 or CO per pound 
of fuel burned, computed from the analysis of the fuel and of the exhaust gases. 

Having determined the volume V 2 of the exhaust gases, formula ( B) may be used in com¬ 
puting the temperature, in which case T will represent the temperature ot the exhaust gases, 
as in the second method, P the pressure of the exhaust, and V the volume of the exhaust 
gases V 2 discharged per stroke, added to the volume of the gases retained in the cylinder at 
the end of the stroke. 

The value of R given in equation (Z>) is approximate, on account of the fact that the 
percentage of N should be that due to the air alone, and not that due to the aii in addition 
to that contained in the fuel gas. Where extreme accuracy is desired, the value found for R 
may be used to determine the percentage of N which in the analysis of the exhaust gases is 
due to the N in the fuel gas, and this value may be subtracted from the total N shown by 
the analysis of the fuel gases, in order to obtain the correct value of N to be used in equation 
(D). 





652 


APPENDIX 


Table 150 

DATA AND RESULTS OF TEST OF GAS OR OIL ENGINE 

Arranged according to the Complete Form advised by the Engine Test Committee, American Society of 

Mechanical Engineers. Code of 1902. 

1. Made by. of . 

on engine located at. 

to determine. 


2. Date of trial. 

3. Type of engine, whether oil or gas.. 

4. Class of engine (mill, marine, motor for vehicle, pumping, or other) 


5. Number of revolutions for one cycle, and class of cycle 


6. Method of ignition 


7. Name of builders.. 

8. Gas or oil used.. 

(a) Specific gravity. deg. Fahr. 

( b ) Burning-point. “ 

(c ) Flashing-point. “ 

9. Dimensions of engine: 

1st Cyl. 2d Cyl. 

(a) Class of cylinder (working or for compressing the charge). 

( b ) Vertical or horizontal.i. 

(c) Single- or double-acting. 

( d ) Cylinder dimensions. 

Bore, in. 

Stroke, ft. 

Diameter piston rod, in... 

Diameter tail rod, in.. ..'. 

(e) Compression space or clearance in per cent of volume displaced by piston per 

stroke. 

Head end... 

Crank end. 

Average.. 

(/) Surface in square feet (average). 

Barrel of cylinders. 

Cylinder heads. 

Clearance and ports. 

Ends of piston. 

Piston rod. 

(g) Jacket surfaces or internal surfaces of cylinder heated by jackets, in square feet 

Barrel of cylinder. 

Cylinder heads. 

Clearance and ports. 

(h) Horse-power constant for 1 lb. M.E.P., and one revolution per minute. 

10. Give description of main features of engine and plant, and illustrate with drawings of same given on 
an appended sheet. Describe method of governing. State whether conditions were constant 
throughout the test. 

Total Quantities 


11. Duration of test. hours 

12. Gas or oil consumed. ‘‘cu.ft. or lbs. 

13. Air supplied in cubic feet. _ _ cubic feet 

14. Cooling water supplied to jackets. '' << 

15. Calorific value of gas or oil by calorimeter test, determined by. B.T.U. 


Hourly Quantities 


16. Gas or oil consumed per hour... 

17. Cooling water supplied per hour 


cu.ft. or lbs. 
lbs. 


















































SOME DETAILS FROM PRACTICE 


653 


Table 150—DATA AND RESULTS OF TEST OF GAS OR OIL ENGINE —Continued 

Pressures and Temperatures 


18. Pressure at meter (for gas engine) in inches of water. ins. 

19. Barometric pressure of atmosphere: 

(o) Reading of height of barometer. “ 

(' b ) Reading of temperature of barometer. deg. Fahr. 

(c) Reading of barometer corrected to 32° Fahr. ins. 

20. Temperature of cooling water: 

(a) Inlet. deg. Fahr. 


( b ) Outlet. 

21. Temperature of gas at meter (for gas engine) 


22. Temperature of atmosphere: 

(a) Dry-bulb thermometer. “ 

(b ) Wet-bulb thermometer. “ 

(c) Degree of humidity. per cent 

23. Temperature of exhaust gases. deg. Fahr. 

How determined. 


Data Relating to Heat Measurement 

24. Heat units consumed per hour (lbs. of oil or eu.ft. of gas per hour multiplied by the total 


heat of combustion). B.T.U. 

25. Heat rejected in cooling water: 

(a) Total per hour). 

(i b) In per cent of heat of combustion of the gas or oil consumed. per cent 

26. Sensible heat rejected in exhaust gases above temperature of inlet air: 

(а) Total per hour. B T.U. 

(б) In per cent of heat of combustion of the gas or oil consumed. per cent 

27. Heat lost through incomplete combustion and radiation per hour: 

(a) Total per hour. B.T.U. 

( b) In per cent of heat of combustion of the gas or oil consumed. per cent 


Speed, Etc. 

28. Revolutions per minute. 

29. Average number of explosions per minute. 

How determined. 

30. Variation of speed between no load and full load...; ■ 

31. Fluctuation of speed on changing from no load to full load measured by the increase in 

the revolutions due to the change. 


Indicator Diagrams 


1st Cyl. 


32. Pressure in lbs. per sq.in. above atmosphere: 

(a) Maximum pressure. 

(b) Pressure just before ignition. 

(c) Pressure at end of expansion. 

(d) Exhaust pressure. 

33. Temperatures in deg. F. computed from diagrams: 

(a) Maximum temperature (not necessarily at maximum pressure) 
(5) Just before ignition. 

(c) At end of expansion. 

( d ) During exhaust. 

34. Mean effective pressure in lbs. per sq.in. 


rev. 


rev. 


2d Cyl. 


Power 

35. Power as rated by builders: ^ p 

(a) Indicated horse-power. /, 

( b ) Brake. 

36. Indicated horse-power actually developed: (( 

First cylinder. , t 

Second cylinder. << 

Total.* * . . . . • tt 

37. Brake H.P., electric H.P., or pump H.P., according to the class of engine. ••••••• . 

38. Friction indicated H.P. from diagram, with no load on engme and computed for a^rage 

speed.. npr cpnt 

39. Percentage of indicated H.P. lost in friction. 










































654 


APPENDIX 


Table 150—DATA AND RESULTS OF TEST OF GAS OR OIL ENGINE —Continued 


Standard Efficiency Results 

40. Heat units consumed by the engine per hour: 

(a) Per indicated horse-power. 

(by Per brake horse-power. 

41. Heat units consumed by the engine per minute: 

(a) Per indicated horse-power. 

(5) Per brake horse-power. 

42. Thermal efficiency ratio: 

(a) Per indicated horse-power. 

(b) Per brake horse-power. 


B.T.U. 

tt 


a 


(( 


per cent 
( < 


Miscellaneous Efficiency Results 

43. Cubic feet of gas or lbs. of oil consumed per H.P. per hour: 

(a) Per indicated horse-power. 

(by Per brake horse-power... 


Heat Balance 

44. Quantities given in per cents of the total heat of combustion of the fuel: 

(ay Heat equivalent of indicated horse-power. 

(by Heat rejected in cooling water. 

(c) Heat rejected in exhaust gases and lost through radiation and incomplete combustion 

Sum = 100 

Subdivisions of Item (c): 

(c t ) Heat rejected in exhaust gases. 

(c 2 ) Lost through incomplete combustion.. 

(c 3 ) Lost through radiation, and unaccounted for. 

Sum = Ttem (c) . 


per cent 

i ( 

( { 
t ( 


< ( 


i ( 
< C 


Additional Data 

Add any additional data bearing on the particular objects of the test or relating to the special class 
of service for which the engine is to be used. Also give copies of indicator diagrams nearest the mean 
and the corresponding scales. Where analyses are made of the gas or oil used as fuel, or of the exhaust 
gases, the results may be given in a separate table. 


Table 151 


DATA AND RESULTS OF STANDARD HEAT TEST OF GAS OR OIL ENGINE 


Arranged according to the Short Form advised by the Engine Test Committee, American Society 

of Mechanical Engineers. Code of 1902 J 


1. Made by.. G f 

on engine located at. 

to determine. 

2. Date of trial. 

3. Type and class of engine. 

4. Kind of fuel used. 

(a) Specific gravity. 

(by Burning-point. 

(c) Flashing-point. 

5. Dimensions of engine: 


deg. Fahr. 

i ( 

H 


(ay Class of cylinder (working or for compressing the charge). 

(by Single- or double-acting. 

(c) Cylinder dimensions: 

Bore, in. 

Stroke, ft. 

Diameter piston rod, in. 

(dy Average compression space, or clearance in per cent. 

(e) Horse-power constant for 1 lb. M.E.P. and one revolution per minute 


1st Cyl. 2d Cyl. 








































SOME DETAILS FROM PRACTICE 655 


Table 151—DATA AND RESULTS OF STANDARD HEAT TEST OF GAS OR OIL ENGINE —Continued 


Total Quantities 

6. Duration of test. 

7. Gas or oil consumed. 

8. Cooling water supplied by jackets. 

9. Calorific value of fuel by calorimeter test, determined by.calorimeter 


hours 

cu.ft. or lbs. 
B.T.U. 


Pressures and Temperatures 

10. Pressure at meter (for gas engine) in inches of water. 

11. Barometric pressure of atmosphere: 

(a) Reading of barometer. 

(5) Reading corrected to 32° Fahr. 

12. Temperature of cooling water: 

(a) Inlet. 

( b ) Outlet. 

(c) Degree of humidity. 

13. Temperature of gas at meter (for gas engine) 

14. Temperature of atmosphere: 

(a) Dry-bulb thermometer. 

(b) Wet-bulb thermometer. 

15. Temperature of exhaust gases. 


ins. 

( C 
i ( 


deg. Fahr. 
( ( 

( < 


deg. Fahr. 

( l 
< ( 


. Data Relating to Heat Measurement 

16. Heat units consumed per hour (pounds of oil or cubic feet of gas per hour multiplied by 


the total heat of combustion). B.T.U. 

17. Heat rejected in cooling water per hour. “ 


Speed, Etc. 

18. Revolutions per minute. rev . 

19. Average number of explosions per minute 

Indicator Diagrams 

20. Pressure in lbs. per sq.in. above atmosphere: 

1st Cyl. 2d Cyl 

(a) Maximum pressure 

( b ) Pressure just before ignition 

(c) Pressure at end of expansion 
(c/) Exhaust pressure 

(e) Mean effective pressure 


Power 

21. Indicated horse-power: 

First cylinder- , . H.P. 

Second cylinder. “ 

Total. 

22. Brake horse-power. 

23. Friction horse-power by friction diagrams. “ 

24. Percentage of indicated horse-power lost in friction. per cent 

Standard Efficiency and Other Results 


25. Heat units consumed by the engine per hour: 

(a) Per indicated horse-power. B.T.U. 

( b ) Per brake horse-power. 

26. Pounds of oil or cubic feet of gas consumed per hour: 

(a) Per indicated horse-power. .bs. or cu.ft. 

(b) Per brake horse-power. 


Additional Data 

Add any additional data bearing on the particular objects of the test or relating to the special class 
of service for which the engine is to be used. Also give copies of indicator diagrams nearest the mean, 
and the corresponding scales. 




























656 


APPENDIX 


2. Rules for Testing Gas Producers and Gas Engines 

(Code of the German Society of Engineers 1 ). 

(All metric units have been transposed to English units.) 

The preparation of the following rules for making gas engine and producer tests was under¬ 
taken by a committee appointed from the Verein Deutscher Ingenieure, in collaboration with the 
German Society of Engine Builders, with the view of establishing definite general regulations 
governing such tests. It is desirable, by specifying the important proportions of the examined 
plants and the conditions under which the results were obtained, to insure that these results 
are not only applicable to a single case, but that they have general value. To attain this end 
it is necessary that all data should be given uniformly according to a code of regulations such 
as that here presented. 

The execution of such tests should be intrusted only to persons possessing the required 
expert knowledge and practical experience. These persons must make a trial plan, or schedule, 
appropriate to the individual case in hand, which, in many instances, will not require that all 
of the investigations stipulated in the general code are actually carried out. They must further 
examine the instruments for measuring or recording purposes as to their fitness and must compile 
the results. The following rules, the adoption or selection of which must be left to the soundness 
of judgment of the investigator, are intended to serve as a basis on which to proceed. 

GENERAL REGULATIONS 

Object of Investigation. 

1. The object of a test made on a producer-gas plant may be to determine: 

(a) The quantity, composition, and calorific value of the fuel consumed. 

(b) The quantity, composition, and heat value of the gas produced. 

(c) The degree of efficiency of the producer-gas plant. 

(i d ) The separate heat losses in the plant. 

(e) The quantity of impurities contained in 1 cm. or 1 cu.ft. of gas (dust, tar, sulphur, 
etc.). 

(/) The moisture contents of the gas. 

(g) The water consumption of the producer-gas plant, either total or in the separate parts. 

(h) The mechanical work required for operating the plant, including apparatus. 

( i ) The duration of time required for starting. 

(k) The stand-by losses during intervals of shutting down day or night. 

2. The object of a test made on an internal-combustion (gas) engine may be to determine: 

(a) The indicated capacity and the effective output. 

(b) The mechanical efficiency. 

(c) The fuel consumption and the heat consumption per horse-power hour. 

(d) The consumption of lubricants, separately for cylinder and engine. 

(e) The consumption of water and the heat conducted to the cooling water. 

(/) The fluctuations in number of revolutions. 

(g) The composition of exhaust gases. 

NUMBER AND DURATION OF TESTS. 

Admissible Fluctuations. 

3. The number and duration of trials are determined by the purpose of the test as well 
as by a consideration of the conditions of installation and operation, and must be settled and 
previously arranged according to paragraphs 4 to 8. For trials of special importance the 
results of which are decisive for acceptance tests, for penalties or for premiums, this item 
deserves special consideration. 

4. Acceptance tests should be made if possible immediately after a plant has been put into 
actual operation; the manufacturers, however, must be granted a reasonable time for making 


1 Mainly from F. E. Junge’s translation in Power, Feb., 1907. 



SOME DETAILS FROM PRACTICE 


657 


preliminary trials of their own and for carrying out alterations or improvements when necessary. 
The length of this time and other conditions are best agreed upon when drawing up the delivery 
contract. 

5. In order to be able to get acquainted with the operation of the plant that is to be 
tested, to find time for examining the testing devices employed, and to break in observers and 
assistants, it is desirable that preliminary trials be allowed. 

6. If the fuel consumption in gas producers is to be determined, the trial run must be 
extended over at least eight hours under constant conditions and without interruptions. 

7. For determining the consumption of liquid or gaseous fuel and provided the conditions 

are constant, it is sufficient for the higher loads to extend measurements over an hour, while for 

finding the consumption at the lower loads, measurements of even shorter duration are sufficient. 
To ascertain the constancy of the conditions the temperature of the outflowing cooling water 

must be read from time to time. These rules as to the duration of the tests are made with 

the provision that no interruption or disturbance of the trial takes place, and that intermediate 
readings show only slightly varying values for the consumption. 

8. If only the mechanical efficiency of an engine is to be determined, trials of short duration 
under constant conditions are sufficient; but at least ten sets of indicator cards should be taken. 

9. For investigations of special importance at least two tests should be made, one after the 
other. They should be accepted only if no interruptions occurred and if the results show no 
greater deviations than those due to unavoidable errors of observation. The mean of the two 
results is to be taken as the final result. 

10. The extent to which the capacity and the consumption of gas may differ from the guarantee 
or contract figures, without justifying a claim of breach of contract, is to be clearly stated 
before the tests (either in the original contract or in the schedule of tests). When no other 
agreement has been previously arrived at, the capacity guarantee is regarded as fulfilled if the 
figure obtained in the test is not more than 5% below the value on which the guarantee was 
based. This margin, however, is allowable only for the maximum output which was promised 
beyond the guaranteed continuous output. The latter must be rendered by the engine under 
all circumstances. 

The consumption of fuel and water as determined on test should not exceed the guaranteed 
figures by more than 5% even if, during the trial, the engine load fluctuated somewhat from 
the load upon which the guarantee was based, provided that fluctuation do not exceed an 
average of ±5% of such load, or a maximum of ±15%. 

Since it is often impossible when making tests to have the internal combustion engine work 
at exactly the effective (horse-power) capacity on which the guarantee agreed upon in the 
contract is based, it is recommended that the agreement shall specify the expected fuel consump¬ 
tion for higher and lower outputs. The same provision is preferably made also with gas 
producers. 

UNITS OF MEASUREMENT AND DESIGNATIONS 

11. When giving pressure data it must be stated whether absolute pressures or gauge 
pressures above or below the atmospheric are meant. Absolute pressure equals atmospheric 
pressure plus gauge pressure. 

12. All temperature and heat measurements refer to the Fahrenheit scale. 

13. The mechanical equivalent of heat is taken at 778 ft.-lbs. 

14. The calorific value of a fuel is to be taken as its lower heating value; that is, the heat 
which is liberated by the complete combustion of the fuel when the burnt products are cooled 
down to the original (room) temperature at constant pressure, it being assumed, however, that 
the water of combustion and the moisture contained in the fuel remain vaporized. The calorific 
value must be based on the unit quantity or weight of original fuel, without deducting ash, 
moisture, etc., and is to be expressed in heat units. For both solid and liquid fuels the unit 

of weight is the pound. , , 

The heat value of gaseous fuels is based on 1 cu.ft. at 32° F., and /60 mm. barometer 
pressure, or must be expressed in thermal units as “effective ’ heat value, that is, reduced to 
1 cu ft of actual gas used. If not specially stated, it is always understood that the heat value 
recorded is that of gas at 32° F. and 760 mm. barometer pressure. „ 

In this country the general standard so far recommended seems to indicate for standard gas 
a temperature of 60° F., and a pressure of 14.7 lbs. per sq.in., corresponding to the usual 

atmospheric pressure. 


658 


APPENDIX 


15. The efficiency of a gas-producer plant is the ratio of the latent heat contained in the 

gas as produced to the heat of combustion of the total weight of fuel consumed in the plant, 
both items being computed from the lower heating value. In producer-gas plants having a 

separately fired steam boiler, it is advisable also to determine the ratio of the heat which is 
chemically bound in the producer gas to the heat equivalent of that portion of the fuel which 
is consumed in the producer proper for making such gas. 

16. The unit of measurement used for the power or work output of an internal-combustion 

engine is the horse-power equal to 33 000 ft.-lbs. per minute. It must be clearly stated whether 
the indicated power, or the useful or available power, is meant. If not otherwise designated it 
is understood that the figures refer to the useful or available output. 

17. The indicated power of the engine, or the indicated work, is the difference between the 

total power developed or work done, and the indicated power, or work, which is consumed 

within the engine; in short, the difference between the positive and the negative indicated power 
or work. 

Note. This is the provision which caused considerable discussion among gas-engine experts some 
time ago. It means, as it stands, that in a 4-cycle machine, the indicated horse-power is that determined 
from the work diagram minus the work shown by the lower loop diagram; and, in a 2-cycle engine, 
the total indicated horse-power, as determined from the diagram of the power cylinder minus the pump 
work, is considered as the indicated horse-power. This view is undoubtedly correct when the mechanical 
efficiency of the engine itself as a machine is to be determined. 

The power required at “no load” is the power indicated when no useful work is rendered by 
the engine. 

18. Mechanical efficiency is the ratio of the useful power to the indicated power of the 
engine. 

19. All consumption figures should be reduced to the hour basis, and if they are to be 
compared with the output of the engine they must be based on one horse-power hour. If not 
otherwise agreed ujion, these data refer to the useful or available output at full load. 


EXECUTION OF TESTS 

20. If the quantity of gas made in a producer or the weight of fuel consumed in an engine 
is to be measured, then all pipes or ducts which are not used in the test must be cut off from 
the piping which leads to the producer and engine that are to be tested. This is best done 
by means of blind flanges. The active ducts, pipes, gas holders, etc., must be examined with 
regard to leakage and made tight if necessary. Unavoidable losses due to leakage must be 
determined. This holds especially for masonry gas mains. 

FUEL CONSUMPTION OF A GAS-PRODUCER PLANT 

21. The kind, number, and duration of tests must be agreed upon according to the general 
rules laid down in paragraphs 1 to 10. 

22. The constructive features and the operative conditions of gas-producer plants must be 
described and illustrated in the report by drawings, so far as this is necessary, to arrive at a 
clear understanding of the manner of working and of the results obtained. 

23. Before making the test the plant should be examined as to whether or not it is in good 
working order. 

24. The quantity of fuel consumed in the gas producer is determined by taking the weight 
of the fuel which is charged into the producer during the trials in order that the producer may 
contain at the end of the test exactly the same amount of heat, either liberated or chemically 
bound in the fuel, that it contained when starting the test. To meet this requirement it is not 
sufficient that the depth of the fuel bed be the same at the end as it was at the beginning; 
it must also be taken into consideration what influence the ash and the slag left in the producer, 
the location of the incandescent zone, the formation of fissures and cavities, the closeness or 
density of the producer charge, and the chemical composition of the burning fuel particles 
exercise on the heat contents of the producer. 

In order to comply with this requirement the following rules should be followed: 

25. When starting the test the plant should be in the condition of stability or normal 
working condition, if possible. This means that after a period of shut-down for cleaning or 


SOME DETAILS FROM PRACTICE 


659 


repairs it should be in active operation for one or more days, running on fuel of the same 
characteristics and size, with the same depth of fuel bed, the same skill of attendance as regards 
the charging or feeding of fresh fuel and the removing of slag, and under the same load 
conditions that will obtain during the test. 

26. During the trial the producer should be charged and poked as nearly in accordance with 
the requirements for attendance as possible. The level of fuel charged must be the same at the 
beginning and at the end of the tests and should be kept constant during the trial. About half 
an hour before starting and before stopping a test, the slag and ashes should be removed. 

If it is impossible to take out the ashes during the operation of the producer, the plant 
must be shut down immediately after stopping the test, the ashes must be taken out at once 
and the producer refilled up to the same level that existed when starting the test. The weight 
of fuel used for this purpose must be added to the consumption. 

27. The fuel consumed during the trial must be weighed, also the fuel which has not been 
burned and remains useful; that is, that portion which drops down from above the grate while 
raking out the ashes, and that which is culled out form the ashes as unburned. The weight of 
the former may be deducted from the consumption, but not the amount which is taken out 
from the ashes, nor the coal dust which accumulates in the scrubbers and in the flues between 
the producer and the engine. 

28. -To be able to determine the quantity of ash and slag produced during the trial, the 
ash box must be emptied before the test. If this is not possible, as when an inclined grate is 
used, the refuse in the ash box must be equalized before and after the run. 

29. The stand-by losses during intervals of shutting down at day and night must be 
determined. 

30. In order to get a representative sample of the solid fuel, the following course may be 
pursued: Of every carload, basket, or other measure of fuel, put a shovelful into a covered 
receptacle. Immediately after the test is over, the contents of the receptacle should be broken, 
mixed, spread, and quartered by drawing the two diagonals of a square. The two opposite 
quarters are rejected, the two others broken up finer, mixed, and quartered, and the two opposite 
quarters rejected. This is continued until a sample of some 10 to 20 lbs. remains, which is 
preserved in well closed receptacles for analysis. In addition to this a number of other samples 
must be put away in air tight receptacles for use in determining the contents of moisture in 
the fuel. 

31. The composition of the fuel shall be determined by chemical analysis. Its contents in 
carbon (C), hydrogen (H), oxygen (0), sulphur (S), ash (A), and water (W) must be given in 
percentage of weight referred to the original fuel. The contents, in the fuel, of nitrogen (N) 
can be disregarded. The behavior of the fuel when being heated should be determined by a 
coking test. 

32. The calorific value of the fuel must be determined by calorimetric analysis. An approxi¬ 
mate determination of the heating value can be made on the basis of the chemical analysis by 
employing Du Long’s formula: 

Heating value = 145 C+522.3 — 

in which C, O, H, S, and W are expressed in weight per cent. 

TESTIXG AN INTERNAL-COMBUSTION ENGINE 

33. Kind, number, and duration of trials to be agreed upon according to the general regula¬ 
tions Nos. 1 to 8. 

34. The constructive features and operative conditions of the engine must be so illustrated 
in the report as to enable one to form a correct idea of the manner of working and of the 
results of operation. Especially important are the type and capacity of engine, diameter of 
cylinder and piston-rod, piston stroke, contents of clearance space, and other essential dimensions; 
the normal rate of revolution and the admissible fluctuations; kind and heat value of fuel for 
which the engine is intended. The diameter of the cylinder and the stroke should be actually 
measured if this is possible. 

The contents of the compression space are preferably determined by filling with water. If it 
is impossible to state the cubical contents of the compression space, then the compression 


O 


+ 40 S-9.66 W, 


660 


APPENDIX 


pressure at full load should at least be given. This is done by taking an indicator card while 
the ignition is interrupted. 

35. Before making the test the engine must be examined internally and externally as to 
whether or not it is in good working order. 

36. The number of revolutions of the engines should be determined by a continuous speed 
counter, the records of which must be noted at certain intervals, and must be checked or 
corrected. from time to time by direct readings. If the speed conditions of the engine are to 
be investigated it is essential to determine the following items: 

(a) The number of revolutions under constant conditions at maximum load and at no load. 

( b ) The fluctuations in speed at constant load. 

(c) The temporary change in the number of turns when the load is suddenly decreased or 
increased from a given constant load by a prescribed amount. These determinations can be 
executed with apparatus of the character of the Horn tachograph. The fluctuations of speed 
during the performance of one engine cycle above and below the mean value, expressed in parts 
of the latter, should be determined by calculation unless otherwise provided. 

The coefficient of fly-wheel regulation is 

Amax -Nmi, n \ 

_ Amax + Amin/ 

2 

where A = number of revolutions. 

37. The useful output can be determined either by brake test or by electrical measurement. 

The dimensions and weight of the brake should be determined before the trial. 

The electrical measurements can be made on a generator directly coupled to the gas engine. 
The useful work is computed from the output of the dynamo. The efficiency of the generator 
should be determined by one of the methods as laid down in the “Rules for Judging and 
Testing Electrical Machinery and Transformers,” published by the Association of German Electrical 
Engineers. If the efficiency is found approximately by measuring the determinable losses, then 
an adequate amount (say 2% of the full load output) must be allowed for losses not accounted 
for. 

The apparatus with which the electrical measurements are executed must be calibrated 
before and if possible also after the test. 

Whether anything besides the 2% above allowed should be credited to the gas engine for 
increased bearing friction and windage of the generator, must be settled in each individual case. 

Whether, in case the useful output can neither be determined by brake test or by electrical 
measurements, the code provision for testing steam engines can be admitted as correct for gas 
engines, namely, to designate the useful output as the difference between the indicated work at 
any load and the indicated work at no load, cannot be settled at the present state of develop¬ 
ment, since results of accurate investigations are not yet available. 

38. Indicators must be connected immediately to the combustion chamber without employing 
long piping with sharp bends, and one indicator must be provided for every combustion chamber. 
For this purpose each compression chamber must have an opening of f or 1 in. Whitworth 
thread. The same holds true for pump cylinders. 

The indicators and their springs must be calibrated before and after the test according to 
the accepted standards. 

39. During the test, cards should be taken quite frequently from every combustion chamber 
and from the pump cylinders. The cards should be designated by numbers, and the time when 
each card was taken, the scale of springs used and the number of single cards obtained must 
be recorded on the cards. At least five diagrams should be taken on one card successively. 
From time to time diagrams indicated with a weak spring should be taken from the combustion 
chambers. 

The indicated work at no load should be determined immediately after stopping the main 
test and while the engine is still warmed up ready for operation. Care must be taken that the 
no load cards are not taken during an acceleration or during a retardation period of the fly¬ 
wheel. 


^ _A max 


-Nr, 


A max + A E 





SOME DETAILS FROM PRACTICE 


661 


ANALYSIS OF THE GAS GENERATED IN A PRODUCER GAS PLANT OR CONSUMED IN AN INTERNAL- 

COMBUSTION ENGINE, OR OF THE LIQUID FUEL USED 

40. The samples for the chemical analysis of the gas must be taken during the trial at 
regular intervals and as frequently as possible. 

They must be either analyzed on the spot or preserved in glass tubes closed by melting the 
ends. The analysis is to determine, in per cent of volume, the contents of the gas in carbon 
monoxide (CO), carbon dioxide (CO,), hydrogen (H,), marsh gas (CH 4 ), heavy hvdrocarbons and 
oxygen (0 2 ). 

In addition it is recommended to determine the contents of sulphur. The gas samples 
should be taken from the gas main between the cleaning apparatus and the engine. 

41. The heat value of the gas should be determined as often as possible by calorimetric 
analysis, and the burner of the calorimeter should be fed from the gas main without interruption. 
In suction producer plants this can be done by means of a gas pump drawing from the main. 
If conditions should make it necessary that a sample be taken from the pipe while the calorim¬ 
eter is shut off, such sample to be later transferred to and burned in the calorimeter, then 
the quantity of gas so taken should not be less than 300 liters (10 eu.ft.), in order that the 
calorimeter may at first be brought into the condition of stability as regards the water of 
combustion, and in order that at least 100 liters (3.5 eu.ft.) remain available for two successive 
analyses. The suction pump, the gas holder and the piping must be made tight with special 
care when making a calorimeter analysis of suction gas. 

42. The gas meter of the calorimeter in which the heat value of the gas is determined must 
be calibrated. For determining the temperatures of the calorimeter water, only thermometers 
with calibration certificates or others compared with such should be used. The scales must be 
divided at least into tenths of a degree. 

On the basis of the chemical analysis the heating value per standard cubic foot of gases 
which do not contain heavy hydrocarbons can be computed from the following formula, if a 
calorimetric analysis cannot be made 


Heating value = 3.42 CO 4-2.97 H 3 + 9.52 CH 4 


where CO, H 2 and CH 4 are expressed in volume per cent. 

43. The quantity of gas produced or consumed should be measured by means of a gas 
holder or a gas meter. The cross-sectional area of the holder should be determined by measure¬ 
ment of its circumference at several places. Consumption tests with the gas holder shall not 
be made while the latter is exposed to the sun. 

44. The gas meter must be calibrated and set level; it must be so filled that the water 
level corresponds to the normal filling existing during calibration. Between the gas meter and 
the engine a pressure regulator must be installed or a large suction space provided so that the 
water level shows only small pulsations during the pressure fluctuations. 

45. At intervals corresponding to the duration of test the following readings should be taken: 
Position of the bell of the gas holder at three places or the records shown by the gas meter; 
the pressure in the bell or in the gas meter; the temperature of the gas when entering and 
when leaving the gas holder or the gas meter and before reaching the engine; the barometric 
pressure. 

46. If the temperature of the gas is different when measuring the consumption than when 
measuring the heat value, the computation must also take into account the increase of volume 
which is due to the moisture contents of the gas at higher temperatures. 

47. The consumption of liquid fuel must be determined either by weight or by measuring its 
volume. For determining heat value, composition, and specific weight of the fuel one representa¬ 
tive average sample is sufficient. 

48. When measuring the fuel consumption of internal-combustion engines, the consumption 
of lubricating oil for the cylinder should be determined at the same time. 

49. If the consumption at low loads of a double-acting tandem or twin engine is to be 
determined, it is not allowable to shut off the gas from one or more ends of the cylinders, 
provided that no other arrangements have been previously agreed upon and are mentioned in 
the report, or that the governor acts automatically in the way described. 


662 


APPENDIX 


EXPLANATIONS TO VARIOUS ARTICLES OF THE CODE 

The main code is followed by a number of explanations from which the following extracts 
are taken. The figures refer to the paragraphs of the above code. 

1 and 2. In most cases only one or two of the objects of test mentioned are taken into 
account in any given trial. If in any exceptional case the object of the test should not be any 
of those mentioned, it should be a simple matter to adapt the rules given. 

Under 2 (c) the term horse-power hour is used. It is essential that in any given trial this 
term be more closely defined, as horse-power may mean indicated brake, or even horse-power 
developed by pumps. 

4. It is extremely desirable that the contract state the time allowed the manufacturer for 
adjustment and trial runs, because his own interests may make him call sometimes for a long, 
sometimes for a short period. In the case of a small engine, more or less a commercial stock 
machine, he may wish to have the period as short as possible, and this the buyer may agree to 
without danger of loss to himself. If, however, the machine is of a special type, or one provided 
with special attachments, it is but a matter of justice to allow the manufacturer a reasonable 
time in which to break in the engine and to give him an opportunity to correct any imper¬ 
fections that may appear. It is to the interest of the buyer to grant such a period in order 
to become familiar with the machine before taking over the entire responsibility of operating it. 
It is also true that many faults appear only after some weeks of operation. 

On the other hand, too long a period of adjustment is in many cases not acceptable to the 
buyer, because any extended work of improvement usually seriously hampers operation; and 
because in many cases he desires an operative machine, which no longer requires the care of the 
manufacturer, as soon as possible. 

It frequently happens that no acceptance test is agreed upon. In such cases it sometimes 
happens that the buyer comes back upon the manufacturer for faults wdiich did not develop 
until the machine had been in operation some time. 

If the manufacturer then agrees to an investigation or a test, a sufficient period should be 
given him to make any investigation he sees fit or to correct any imperfections that may have 
appeared before the decisive trial or investigation is made. This sometimes leads to a simple 
settlement of the matter in that the manufacturer discovers that ignorance or carelessness on the 
part of the operator have caused the imperfections complained of. The granting of such a period 
also guards the buyer against any later claim of the manufacturer that during the trial the 
machine was not in the condition in which he delivered it. 

5. Preliminary tests are always desirable, but not absolutely necessary. The cost of any 
kind of investigation is usually quite high and of course the cost increases directly with the time. 
The expert called will therefore make such tests when they seem to him essential. But the 
manufacturer should have the right to call for the time necessary for such trials if he is to 
present the machine in its best condition. 

6. It cannot be denied that eight hours is a rather short time, because it is extremely 
difficult to determine whether the producer is in the same condition at the end as at the 
beginning of the test, and because this uncertainty may lead to large errors. On the other hand, 
it is unquestionable that in many cases a longer time would call forth so many difficulties in 
operation that eight hours would seem the necessary limit. 

The rule is mainly framed to prevent trials of so short duration that serious errors can 
hardly be avoided, but it leaves it to the judgment of the experimenter whether to make the 
tests longer than eight hours where it seems desirable and is possible to do so. 

7. Intermediate readings are recommended without qualification, since they form the best 
criterion of the constancy of conditions. With liquid or gas fuels of constant composition, 
individual readings every five minutes apart sometimes show no variation for hours at a time. 
In such a case it is useless to extend the time of the trial. 

8. In determining the mechanical efficiency of an engine it should not be forgotten that, 
although the average load may be constant, there may be speed variations due to the inevitable 
inequality of the power impulses, so that during some cycles work is done in accelerating the 
fly-wheel, while during others the fly-wheel by retardation gives up some of its kinetic energy. 

To minimize any error that this may introduce into the determination of the mechanical 
efficiency, at least ten indicator diagrams should be taken. 

If the conditions are otherwise constant, however, it is not necessary to spread these 
diagrams over any considerable period of time. 


SOME DETAILS FROM PRACTICE 


663 


It is self-evident that during the time of taking the diagrams the supply of lubricating oil 
must not be increased. 

Changes in the mechanical efficiency of the engine, .as for instance those due to fouling, 
cannot be detected with certainty even^by a long test period; they become noticeable usually 
only after a period of operation extending over two weeks. The determination of the mechanical 
efficiency of an engine, after constant conditions of operation are attained, therefore only applies 
to the engine in its then existing state or condition. 

The number of diagrams to be taken on one card cannot be definitely stated. On account 
of variation in the diagrams, which is less at high than at low loads, care should be had not to 
take too few. On the other hand, it is useless to take more than can be clearly distinguished. 
The running together of a larger number of diagrams only makes their evaluation more uncertain. 

10. In consideration of unavoidable errors of observation, possible errors of the instruments 

used, etc., it is meet and usual to allow a certain margin between the figures found on trial and 

those guaranteed. In the steam engine code 5% is allowed for this, and it seems reasonable to 
assume the same figure in this case. Only in one point, in the guaranteed normal capacity, 
does the gas engine call for an exception. 

A given steam engine gives its most economical results at a certain cut-off, but a higher 
capacity can always be obtained at the expense of a little economy, that is, a buyer is certain 

that even a machine slightly too small will give him sufficient capacity. A gas engine, on the 

contrary, works with the greatest economy at its maximum load. It is to the interest of the 
buyer, therefore, to get an engine exactly suited to his needs and not to choose it too large. It 
is possible for the same reason that any engine, if lacking slightly in guaranteed capacity, may 
become absolutely useless to the buyer. For these reasons it was thought advisable not to 
grant the manufacturer any leeway whatever as regards guaranteed capacity. 

It is clear, therefore, that the manufacturer must take upon himself any possible inaccuracies 
in the measurements, unless he can show them up and demand a new trial. For that reason it 
is well for him to make his guarantee a little on the safe side of what he knows his engine is 
capable of developing. On the other hand, there is no harm done to the interest of the buyer 
if the manufacturer underrates the normal capacity of his machine, because the former will 
always call for an engine of a certain normal capacity to suit his needs. If he fails to do this, 
but places his dependence in the guaranteed maximum capacity, he is open to the charge of 
carelessness. 

Since during acceptance tests it is often not possible to keep the load quite constant, it 
became necessary, following the steam engine code, to allow a certain amount of variation, 
within which no just cause could be found for objection to the trial. There are cases'where the 
variations occurring are much greater, as when a gas engine is used for driving a roll train. 
But no one set of rules can possibly take into account all such extreme cases, and in such 
instances the contract should contain the necessary agreements to make atry test clear and free 

from subsequent objections. . . 

The wish has been expressed from several quarters, that the rules should contarn a definrtron 
of the term “normal capacity.” On account of the peculiarity of the gas engine above discussed, 
this is not quite feasible. But the term “maximum continuous capacity” perhaps defines most 

nearly what is intended in most cases. . , 

14. It is sometimes the case that the heating value of the standard cubrc foot, that rs, 
reduced to 32° F. and 760 mm. barometer, is so greatly different from the actual value of the 
gas as used that any contract which contains only the heating value of the gas stated on that 
basis does not convey much meaning to the non-technical buyer. If for instance a given gas 
has a heating value of 135 B.T.U. per standard cu.ft., its effective heating value at a high 
altitude and in a warm climate, say as 68° and 620 mm barometer, will only be about 100 
B T U per cu ft To obviate any misunderstanding, it should be cleai ly stated that, when the 
effective heating’value of the gas is not definitely specified, the heating value at 32° F. and 

760 mm. baromuter is meant. , 1A 

19. By full load is meant the normal capacity, as per paragraph 10. 

23 For acceptance tests, and all other tests which are intended to decide any disagreements 
between manufacturer and buyer, such examination should be earned out m the presence and 
with the aid of the former, as already mentioned under paragraph 4 - 

24-26 In all gas producer tests it is hardly possible with certainty to have all conditions 
exactly'the same at the end as at the beginning. But since any difference in the beginning 
and end conditions may lead to considerable error, which can only be equalised by excessive 




664 APPENDIX 

length of test, the rules are intended to operate to the end that such errors are not in any way 
magnified by the method of test. Hence the detailed statement in the regulations. 

27. Since in actual operation the .fuel in the ash or the coal dust in the gas mains are 
hardly ever utilized, no correction should be made for these on any trial. In order to prevent, 
however, the results from being influenced by insufficient cleaning of the producer, any fuel 
which falls out from above the grate during the cleaning period may be subtracted from the 
amount charged. 

35. See explanation under 23. 

37. A brake test of a large engine is in some 'nstances not possible, and in any case a 
matter of considerable cost. In many cases, however, the larger gas engines are either direct 
connected to a generator or to some other power consumer, as a blowing cylinder. In the 
former case electrical measurements, from which the effective horse-power may be determined, 
are easily made. In the latter case the capacity guarantee will in most instances be based upon 
the performance of the power consumer, as for example the air compressed by the blowing 
cylinder. Outside of engines of this type, however, there still remain many cases in which it 
would be of the utmost value to have some means of determining the effective capacity, and 
it should not be forgotten that, even in the case of medium-sized machines, a braking of the 
engine at the place of erection is often, on account of local restrictions, very difficult. The 
problem has been solved for steam engines by assuming that the difference between the indicated 
horse-power at any load and the indicated horse-power at no load is the effective or useful 
horse-power. It is quite possible that in many cases this is not quite correct, but the method 
is very generally accepted and followed. 

On account of the great overload capacity of the steam engine, a small error in this respect 
does not mean a great deal. But the case of the gas engine is quite different. The data on 
hand does not warrant the application of the same method to the gas engine, and the conse¬ 
quences of an erroneous conclusion are much more serious on account of the lack of overload 
capacity. 

For these reasons one is compelled in some cases to omit the determination of the effective 
capacity altogether and to be content with the determination of the indicated power only. It is 
recommended in such cases that the mechanical efficiency be not assumed too high and that any 
guarantees regarding fuel, etc., also be based upon the indicated horse-power. 

It is sometimes possible to brake an engine on the test floor of the factory. The mechanical 
efficiency may thus be previously determined when it is known that no brake test can be made 
in the final place of erection. 

39. The number of diagrams to be taken during any given test cannot be definitely specified. 
Much depends upon the length of test, and the decision may be left to the judgment of the 
experimenter. 

It is, however, always recommended that a bundle of diagrams, instead of only one, be 
taken on every card. Thus a series of diagrams are obtained, while, if only a single diagram 
is taken, it is possible to hit upon the same diagram in the series a number of times. (See under 
extract 8.) 

The work of fluid friction, that is, the lower-loop diagram, cannot be determined with 
certainty from the full indicator cards. It is best for that reason to ignore the loop when 
determining the positive work and to find the negative work from special weak spring diagrams. 

48. The measurements of the quantity of lubricating oil used is of importance in smaller 
engines, because the fuel consumption can be favorably influenced by a copious supply of the 
lubricant. 

49. If under low loads, only one end of the cylinder is allowed to work, the fuel consump¬ 
tion would be much lower. But since this is not generally done in operation, the results would 
be erroneous. If, however, the governor during operation shuts off the individual cylinders or 
cylinder ends, as the load drops, this is of course also permissible during a test. 


INDEX 


Absolute temperature, 566 
Acetylene, 526 
Adiabatic, 571 
Adiabatic changes, 570 
Adiabatic, false, 573 
Adjustment of valve gearing, 214 
Air pipe line, 302 
Alberger gas engine, 486 
Alcohol, 530 
Alcohol engines: 

Deutz, test of, 333 

Allis-Cha liners (N urn berg) engines, 475 
Allowable fiber stresses (see under name of part 
considered) 

Allowable fiber stresses in materials, 70 

Allowable load on foundations, 258 

American Diesel engines, 429 

American Diesel engines, table of dimensions, 432 

American Diesel engines, tests of, 435 

American Engines: 

Alberger gas engine, 486 

Allis-Cha liners (Niirnberg) engines, 475 

Diesel engine, 429 

DeLa Vergne 2-cycle oil engine, 427 
Hornsby-Akroyd oil engines, 426 
Korting 4-cycle engines, 410 
Korting 2-cycle engines, 410 
Mesta gas engine, 473 
Olds gas engines, 451 
Olds gasoline engines, 457 
Riverside gas engine, 469 
Snow gas engines, 445 
Struthers-Wells gas engine, 490 
Tod gas engine, 481 
Westinghouse horizontal engines, 385 
Westinghouse producer and power plant with 
Loomis-Pettibone producers, 401 
Westinghouse vertical gas engines, 397 
American gas engines, types of, 385 
Avogadro’s Law, 580 
Arrangement of cylinders, 63 
A.S.M.E. Code for testing gas and oil engines, 642 
Atomic heat, 548 


Atomic weight, 580 
Atomizers, 263 
Atomizers, designs of, 265 
Atom, definition of, 580 
Attendance, cost of, 315 
Automobile engines, designs of, 503 

Banki engines, 379 
Banki gas engine, test of, 383 
Banki gasoline engine, test of, 382 
Bearings, main, 113 
Blast-furnace gas, 521 
Blast-furnace gas engines: 

Cockerill, tests of, 369 
Blowing engines: 

Letombe, test of, 377 
Oechelhauser, test of, 351 
Bolts, table standard dimensions, 306 
Bosch magneto, 253 
Boyle’s Law, 568 
Brass unions, dimensions of, 309 
Brown coal gas, 525 
Buildings, cost of maintenance, 318 

Cams, 213 

Cam shaft (see lay shaft), 212 
Capacity coefficients, 4-cycle engines, table of, 81 
Capitaine double producer, 285 
Carburetor, Olds, 460 
Carburetors, 261 
Designs of, 262 
Card factor, 7 
Carnot cycle, 578 
Centrifugal governors, 244 
Centrifugal washers, 289 
Classification of engines, 1 
Cleaning material, cost of, 316 
Closed cycles, 578 
Cockerill engines, 367 

Cockerill engines using blast-furnace gas, tests of, 
369 

Coke-oven gas, 524 

Combustion chamber, volume of, 123 

665 



666 


INDEX 


Combustion, current opinions regarding, 549 
Combustion, definition of, 585 

Combustion and expansion stroke, 4-cycle engine, 34 
Combustion in tbe gas engine, 539 
Combustion in gas engines (older views), 541 
Combustion temperature, 586 
Compagnie Duplex, 374 
Comparison of cycles, thermodynamic, 20 
Comparison of engines with and without crosshead, 
60 

Comparison of 4- and 2-cycle engines, 60 
Comparison of 4- and 2-cycle engines from design 
standpoint, 52 

Comparison of 4- and 2-cycle engines from stand¬ 
point of economy, 53 

Comparison of fuel cost of steam plant and suction- 
gas plant, 360 

Comparison of horizontal and vertical engines, 58 
Comparison of single- and double-acting engines, 61 
Complete expansion engines, 66 
Compounding of gas engines, 66 
Compression pressures for different ratios of expan¬ 
sion and compression, 33 
Compression stroke, 4-cycle, consideration of, 32 
Compression temperature and pressure for different 
ratios of compression in 4-cycle engines, 35 
Compressors for starting engines, 294 
Computations for strength of crank shafts, 163 
Connecting-rods: 

Constructive details of, 191 
Designs of, 188 
Material, 188 

Cooling arrangements, 299 
Cooling ponds, 301 
Cooling tanks, 300 
Cooling towers, 301 
Cooling tower pipe lines, 303 
Cost of attendance, 315 
Cost of cleaning material, 316 
Cost curves, 321 
Cost of erection, 313 
Cost of lubrication, 316 
Cost of maintenance of buildings, 318 
Cost of maintenance of engines and auxiliaries, 
318 

Costs of fuel, 314 

Costs of fuel, comparison of in steam an suction- 
gas plant, 360 

Costs, table of, for illuminating- and suction-gas 
installations, 319 
Crank arm, checks for strength: 

(a) Dead center position, 167 

(5) Maximum turning moment, 171 
Crank pin, design of, 170 
Crank-shaft, strength computations, 163 

(a) For dead center position, 164 

(6) For maximum turning moment, 170 
■Crank-shafts (multi-throw): 

Design of, 181 


Crank-shafts (single-throw): 

Allowable fiber stress, 155 
Design of, 184 

General kinematic and stress relations in, 155 
Material of, 155 

Crossley brown coal gas producer, 285 
Cycle, Carnot, 578 
Cycle, polytropic, 578 
Cycles: 

Closed, 578 

Constant pressure, practical considerations of, 19 
Constant pressure, table of thermal efficiencies, 
19 

Constant pressure, thermodynamic examination 
of, 16 

Constant volume, practical considerations of, 13 
Constant volume, relations between compression 
and efficiency, 14 

Constant Volume, table of thermal efficiencies, 13 
Constant volume, thermodynamic examination 
of, 11 
Open, 578 

Cylinder covers and heads: 

Allowable stress in, 131 
Constructive details of, 134 . 

Designs of, 132 

Fundamental requirements for, 131 
Material, 131 

Cylinder diameter, determination of according to ah 
required for combustion, 73 
Cylinder diameter, determination of, thermody¬ 
namically, 72 
Cylinder head studs: 

Allowable stress in, 139 
Material, 139 
Number of, 139 
Cylinders: 

Allowable fiber stress in, 120 
Constructive details of, 123 
Designs of, 122 
Material, 120 

De La Vergne 2-cycle oil engines, tests of, 427 
De La Vergne vertical 2-cycle engines, 421 
De La Vergne vertical 2-cycle engines, table of 
dimensions, 421 
Density, 564 
Depreciation, 316 
Deschamps producer, 284 
Design of pistons, 141 
Deutz, alcohol engine, test of, 333 
Deutz double zone generators, 283 
Deutz gas engines, 324 
Deutz, illuminating-gas engine, test of, 327 
Deutz, petroleum and gasoline engine, test of, 332 
Deutz pressure-gas producers, instructions for opera¬ 
tion of, 630 

Deutz, pressure-producer gas engine, test of, 328 
Deutz, suction-gas installation, test of, 329 
Deutz, suction-gas producers, 274 


INDEX 


667 


Deutz, suction-gas producers, instructions for opera¬ 
tion of, 634 

Diagrams, thermo-chemical, 598 
Diameter of cylinder, determination of according to 
air required for combustion, 73 
Diameter of cylinder, determination of thermo¬ 
dynamically, 72 

Diameter of cylinder, equations for determination of 
for engines using liquid fuel, 75, 79 
Diameter of cylinder, equations for determination of 
for engines using gas fuel, 74, 77 
Diesel engine, American, 429 

Diesel engines, American, table of dimensions, 432 
Diesel engines, American, tests of, 435 
Diesel engines, German, 362 

Diesel engines, German, fuel consumption, table of, 
366 

Diesel engines, German, table of dimensions, 433 
Diesel engines, German, tests of, 363 
Diffusion, 583 

Directions for operation, attendance, etc., of gas 
engines, producer^, etc., 610 
Directions for starting Hornsby-Akroyd engines, 
628 

Dissociation, 583 
Distillation, 592 

Double zone generators, Deutz, 283 
Driving gears, 211 
Dry purifiers, 289 
DuLong’s formula, 591 
Duplex engines, 374 

Economy, comparison of 4- and 2-cycle engines from 
standpoint of, 53 
Efficiency, economic, 9 
Efficiency, indicated thermal, 8 
Efficiency, mechanical, 9 
Efficiency, volumetric, real, definition of, 29 
Efficiency, thermal, definition of, 578 
Efficiency, volumetric, apparent, definition of, 29 
Efficiencies for constant pressure cycle, table of, 19 
Efficiencies for constant volume cycle, table of, 13 
Electric igniters, 252 

Electric ignition apparatus, designs of, 253 
Elements, 581 
Engine beds, 102 

Engine frames, for horizontal engines, 104 
Engine frames, for vertical engines, crank-shaft 
above cylinder, 86 

Engine frames, for vertical engines, crank-shaft 
below cylinder, 90 

Engine, pistons for double-acting, 140 
Engine, pistons for single-acting, 140 
Engines, portable, 493 
.Engines, American: 

Alberger, 486 
Allis-Chalmers, 475 
De La Vergne (oil), 421 
Die el, 429 


Engines, American: 

Hornsby-Akroyd, 416 
Korting 4-cycle, 410 
Korting 2-cycle, 410 
Mesta, 473 
Olds, 451 
Riverside, 469 
Snow, 445 
Struthers-Wells, 490 
Tod, 481 

Westinghouse, 385 
Engines, German, ypes of, 324 
Deutz, 324 
Diesel, 362 
Gnom, 361 
Giildner, 354 
Korting 4-cyclo, 334 
Korting 2-cycle, 334 
Niirnberg, 344 
Oberursel, 361 
Oechelhauser, 345 
Soest, 353 

Engines, other foreign makes: 

Banki, 379 
Cockerill, 367 
Duplex, 374 
Ganz, 379 

Langen & Wolf, 367 
Letombe, 376 
Premier, 377 
Engines, American, 385 
Engines, automobile, 503 
Engines, classification of, 1 

Engines, comparison of horizontal and vertical types, 58 
Engines, comparison of single- and double-acting, 61 
Engines, comparison of, with and without cross-head, 
60 

Engines, complete expansion, 66 
Engines, cost of maintenance, 318 
Engines, gasoline fire engines, Deutz, 497 
Engines, launch, 503 
Entropy, 579 
Entropy diagram, 10 

Entropy diagrams for constant-volume and constant- 
pressure engines, 41 
Erection, cost of, 313 
Exhaust mufflers, 297 
Exhaust pipe line, 303 
Exhaust stroke, 4-cycle engine, 39 
Expansion and combustion stroke, 4-cycle engine, 39 

Fastenings and connections for split flywheels, 235 
Fire engines (gasoline) Deutz, 497 
Flanges, oval, dimensions of, 311 
Flywheel regulation for various cylinder combina¬ 
tions, 230 

Flywheel rim weights, table of, 231 
Flywheels: 

Allowable fiber stress in, 215 


668 


INDEX 


Fly-wheels: 

Constructive details of, 233 

Designs of, 232 

Determination of weight, 215 

Fastenings and connections for split wheels, 235 

Material, 215 

Static computations for, 233 
Foundation construction, 258 
Foundations, 257 
Foundations, material of, 258 
Frames: 

Allowable fiber stress in, 85 
Constructive details of, 87 
Designs of, 86, 104 
For horizontal engines, 104 
Material for, 85 

For vertical engines, crank-shaft above cylinder, 

86 

For vertical engines, crank-shaft below cylinder, 90 

(а) Box frames, design of, 91 

(б) A-Frames, design of, 94 
Friction losses in 4- and 2-cycle engines, 51 
Friction losses in journals, 160 

Fuel costs, comparison of in steam and suction-gas 
plant, 360 

Fuel consumption, Diesel engines, 366 
Fuel consumption, Diesel engines (German), table of 
values, 366 
Fuel costs, 314 
Fuel gases, 514 
Fuel mixtures, 533 
Fuel pumps, 266 
Fuels, 511 

Fuels, standard condition of, 5 

Ganz & Co., 379 
Gas: 

Acetylene, 526 
Blast-furnace, 521 
Brown coal, 525 
Coke oven, 524 
Illuminating, 514 
Lignite, 525 
Natural, 526 
Oil, 516 
Power, 517 
Producer, 517 
Water, 594 
Gas bags, 305 

Gas engines, compounding of, 66 
Gas hammers, Ganz & Co., 385 
Gas meters, 304 
Gas-pipe line, 302 
Gas-pipe lines for power gas, 303 
Gas producer, suction type, Olds, 466 
Gas producers, design of, 285 
Gas producers, designs of: 

(a) Pressure-gas installations, 272 
(5) Suction-gas installations, 273 


Gas receivers, 290 
Gas washers, 288 
Gases, fuel, 514 
Gasification, 592 
Gay-Lussac’s Law, 568 
Gearing, valve, 204 
Gears, driving, 211 

Gears, screw, table of dimensions, 211 
Generators, double-zone type, Deutz, 283 
Generators, power-gas, 269 
German gas engines, types of, 324 
Gnom engines, table of dimensions, 361 
Governing: 

Methods of, 240 

(a) Hit and miss, 240 
(5) Quality, 240 
(c) Quantity, 240 
Governing valve, Tod engine, 486 
Governing valves, Alberger engine, 490 
Governor regulation, coefficient of, 246 
Governor, Tod engine, 487 
Governors: 

Centrifugal, 244 
Construction of, 242 
Design of, 243 

Details of construction of, 244 
Pendulum governors, 242 
Special governors, 245 
Gravity, specific, 564 
Giildner engines, 354 

Giildner illuminating-gas engine, test of, 357 
Giildner suction-gas engine, test of, 357 
Giildner suction-gas installation, 280 
Giildner suction-gas plants, table of dimensions, 355 

Hammers, gas, Ganz & Co., 385 
Heat, atomic, 584 
Heat interchanges in a cycle, 40 
Heat units, definition of, 563 
Heating lamps, designs of, 265 
Heating value, 586 
Heating value, higher, 5 
Heating value, lower, 5 
Heating value of hydrocarbons, 593 
Heat, specific, 2 

Heats, specific, table for perfect gases, 567 
Heats, specific, variation of, 3 
Helical springs, 312 

Hornsby-Akroyd engine, directions for starting, 628 
Hornsby-Akroyd oil engines, 416 
Hornsby-Akroyd oil engines, table of dimensions, 419 
Hornsby-Akroyd oil engines, tests of, 426 
Hot tube igniters, 250 
Hydrocarbons, heating value of, 593 

Igniters, electric, 252 

Igniters, hot tube, 250 

Ignition apparatus, 250 

Ignition system, Westinghouse engines, 395 


INDEX 


669 


Ignition temperature, 586 
Illuminating gas, 514 
Illuminating-gas engines: 

Deutz, test of, 327 
Giildner, test of, 357 
Korting (German), test of, 342 
Indicator diagram, standard, 70 
Inlet mufflers, 296 

Installations, producer, types of, 269 
Instruction book, Olds engine, 611 
Interchanges of heat in a cycle, 40 
Interest and depreciation, 316 
Isobar, 571 
Isothermal, 572 
Isothermal changes, 568 

Jacket wall, design of, 128 
Jackets, 120, 122 
Joule’s equivalent, 2 

Journals, main (see under Main journals) 

Keys, 308 

Keys, table of dimensions, 308 
Korting 4-cycle engines (American), 410 
Korting 4-cycle engines (German), 334 
Korting 4-cycle engines (American), table of dimen¬ 
sions, 409 

Korting 2-cycle engines (American), 410 
Korting 2-cycle engines (American), table of dimen¬ 
sions, 408 

Korting 2-cycle engines (American), tests of, 423 
Korting 2-cycle engines (German), 334 
Korting illuminating-gas engine, (German), test of, 
342 

Korting pressure-gas installations, (German), table 
of dimensions, 341 
Korting pressure-gas producers, 272 
Korting pressure-gas producers, instructions for 
operation of, 633 

Korting pressure producer gas engine (German), 
test of, 343 

Korting suction-gas plant, test of, 344 
Korting, suction-gas producer, 274 

Launch engines, designs of, 503 
Lay shaft, 212 

Lencauchez pressure-gas producers, 273 

Letombe blowing engines, test of, 377 

Letombe engines, 376 

Levers, valve, 214 

Lift of valves, 201 

Lignite gas, 525 

Liquid fuels, 526 

Locomobiles, designs of, 494 

Locomotives, motor, 506 

Loomis-Pettibone producers, 401 

Losses by friction in 4- and 2-cycle engines, 51 

Lubrication, cost of, 316 


Lubrication of: 

Pistons, 147 
Main bearings, 114 
Wrist pins, 147 

Magneto, Bosch, 253 
Main Bearings: 

Constructive details of, 115 
Designs of, 114 
Lubrication of, 114 
For horizontal engines, 113 
For vertical engines, 113 
Main journals, design of, 162 
Maintenance of buildings, cost of, 318 
Maintenance of engines and auxiliaries, cost of, 318 
Mallard & Le Chatelier, 3 

Material for engine parts, (see under name of part) 

Material of foundations, 258 

Mean effective pressure, indicated, 81 

Mean effective pressure, useful, 83 

Mechanical equivalent, 574 

Mesta gas engine, specifications, 474 

Meters, gas, 304 

Mixture, fuel, 533 

Molecule, definition of, 580 

Motor locomotives, 506 

Motor plows, 498 

Motor plows, Ganz, 498 

Motor plows, Oberursel, 500 

Motor vehicles, 501 

Motors, for starting gas engines, 296 

Mufflers, designs of, 298 

Mufflers, inlet, 296 

Mufflers, exhaust, 297 

Multi-cylinder arrangements, 63 

Multi-throw crank-shafts, design of, 181 

Natural gas, 526 
Nozzles, spray, 301 

Niirnberg engines (American), see under Allis- 
Chalmers 

Niirnberg engines (German), 344 
Niirnberg engines (German), table of, 344 
Niirnberg engines (German), tests of, 345 

Oechelhauser blowing engine, test of, 351 
Oechelhauser engines, 345 
Oil engines: 

Deutz, test of, 332 
Diesel (American), 429 
Diesel (German), 362 
Hornsby-Akroyd, 416 
Oil, crude, and its distillates, 526 
Oil gas, 516 
Olds carburetor, 460 
Olds engines, instruction book, 611 
Olds gas engines, 451 
Olds gas engines, table of dimensions, 458 
Olds gas engines, tests of, 466 


670 


INDEX 


Olds gasoline engines, 457 

Olds gasoline engines, table of dimensions, 464 

Olds suction-gas producer, 466 

Open cycles, 578 

Operating results (see under Tests of) 

Operation of gas engines, directions for, 610 
Operation of gas engines and producers, directions 
for, 610 

Packing, methods of packing cylinder heads, 137 
Pedestals (see Foundations) 

Designs of, 257 
Pendulum governors, 242 
Perfect gas, table of constants for, 581 
Pintsch suction-gas plants, 275 
Pipe, purge, 290 

Pipe, wrought iron, dimensions of, 309 
Piping: 

Air pipe line, 302 
Cooling water lines, 303 
Dimensions of pipe lines, 304 
Exhaust-pipe line, 303 
Gas-pipe line, 302 
Power-gas lines, 303 
Pistons: 

Allowable fiber stress in, 140 
Design of, 141 

For double-acting engines, 140 
For single-acting engines, 140 
Material, 140 
Piston rods: 

Design of, 141 

Principal dimensions, determination of, 154 
Piston rings: 

Constructive details of, 148 
Material, 147 
Number of, 149 
Position of, 146 
Spring pressure of, 149 
Table of values for snap rings, 150 
Plows, motor, 498 
Poisson’s Law, 570 
Polytropic cycle, 578 
Ponds, cooling, 301 
Portable engines, 493 
Power gas, 517 

Power-gac producers (see under Gas producers) 
Premier gas engines, 377 
Premier Gas Engine Co., 377 
Premier gas engines, test of, 378 
Pressure-gas installations, Korting (German^ table 
of dimensions, 341 

Pressure, indicated mean effective, 81 
Pressure regulators, 305 
Pressure, useful mean effective, 83 
Pressure-volume curves, examination and construc¬ 
tion of, 571 

Pressures for scavenging 2-cycle engines, 45 
Producer, double, Capitaine, 285 


Producer for brown coal, Crossley, 285 
Producer, Deutz pressure-gas, 272 
Producer, Deutz suction, 274 
Producer, Korting pressure-gas, 272 
Producer, Korting suction, 274 
Producer Lencauchez, 273 
Producer-gas engines: 

Deutz, pressure, test of, 328 
Deutz, suction, test of, 329 
Giildner, suction, test of, 357 
Korting (German), pressure, test of, 343 
Producers for fuels carrying tar, 282 
Producer gas, 517 

Producer installations, types of, 269 

Producers, power-gas, 269 

Pump operations, 2-cycle engines, 42 

Pumps, design of fuel, 267 

Pumps, for liquid fuel, 266 

Purchasing of machinery, specifications for, 637 

Purge pipe, 290 

Purifiers, 298 

Purifiers, dry, 289 

Radiators, surface, 301 
Railways, gas, 506 

Ratio of stroke to diameter and speed of rotation, 67 
Receivers, gas, 290 

Reciprocating parts, table of weights per sq.in. of 
piston face, 157 
Regulation coefficients, 231 

Regulations for installation of alcohol engines for 
agricultural purposes, 641 

Regulations for installation of gasoline engines for 
agricultural purposes, 640 

Regulations concerning installation of internal com¬ 
bustion engines, 637 

Regulations for installation of kerosene engines for 
agricultural purposes, 641 

Regulations for installation of suction-gas power 
plants, 639 

Regulations concerning testing of gas engines and 
gas producers, 642 
Regulators, pressure, 305 

Relation between heat content of a mixture and 
economic efficiency, 22 
Rim weights for flywheels, table of, 231 
Riverside gas-engine, specifications, 469 
Rules for testing gas engines and gas producers 
(German Society of Engineers), 656 

Scavenging phenomena in 2-cycle engines, 43 
Scavenging pressures in 2-cycle engines, 45 
Screens, 306 

Screw gears, table of dimensions, 211 

Screw threads, table of standard dimensions, 306 

Screws, table of allowable loads, 307 

Size of foundation required, 260 

Slaby’s formula, 592 

Snow engines, tests of, 448 


INDEX 


671 


Snow gas engines, 445 

Snow gas engines, table of dimensions, 445 

Societe anon., Letombe, 376 

Soest engines, 353 

Specific heats, 2 

Specific heats, table for perfect gases, 567 
Specific heats, variation of, 3 
Specific gravity, 564 

Specifications for purchasing of machinery, 637 
Specifications, Mesta gas engine, 469 
Specifications, Riverside gas engine, 469 
Spray nozzles, 301 
Springs, helical, 312 
Spring table, 312 
Standard indicator diagram, 70 
Starting apparatus, 290 
Hand crank, 291 
Mechanical, 291 
Starting by compressed air, 292 
Starting by electricity, 295 
Starting by fuel mixture, 295 
Starting engines, compressors for, 294 
Starting motors, 296 

Strength computations, crank-shafts, 163 
Stresses allowable in materials, 70 
Stress relations in single-throw crank-shafts, 155 
Stroke, determination of according to air required 
for combustion, 73 

Stroke, determination of thermodynamically, 72 
Stroke, equation for determination of, for engines 
using gas fuel, 74, 77 

Stroke, equation for determination of, for engines 
using liquid fuel, 75, 79 
Studs, 306 

Studs for cylinder heads (see under Cylinder head 
studs), 139 
Stuffing boxes: 

Design of, 138 

Table of principal dimensions, 139 
Suction-gas installation, Giildner, 280 
Suction-gas plants, Pintsch, 275 
Suction-gas producer, Olds, 466 
Suction pressures, table of, 31 
Suction stroke, 4-cycle, consideration of, 28 
Surface radiators, 301 

Tar-carrying fuels for producers, 282 
Temperature, definition of, 563 
Temperature of combustion, 586 
Temperature of ignition, 586 

Testing of gas engines and gas producers, regulations 
of A.S.M.E., 642 

Testing of gas engines and gas producers, regulations 
of German Society of Engineers, 656 
Tests or operating results of: 

Banki gas engine, 383 
Banki gasoline engine, 382 
Cockerill engines using blast-furnace gas, 369 
De La Vergne 2-cycle oil engine, 427 


Tests or operating results: 

Deutz (German): 

Alcohol engine, 333 
Illuminating-gas engine, 327 
Petroleum and gasoline engine, 332 
Pressure producer-gas engine, 328 
Suction-gas installation, 329 
Diesel (American), 435 
Diesel (German), 363 
Giildner illuminating-gas engine, 357 
Giildner suction-gas engine, 357 
Hornsby-Akroyd oil engine, 426 
Korting (American): 

Two-cycle gas engine, 423 
Four-cycle gas engine, 423 
Korting (German): 

Illuminating-gas engine, 342 
Pressure producer-gas engine, 343 
Suction-gas plant, 344 
Letombe blowing engine, 377 
N urn berg engines (German), 345 
Oechelhauser blowing engine, 351 
Olds gas engine, 466 
Premier gas engine, 378 
Snow gas engine, 448 
Westinghouse horizontal gas engine, 404 
Westinghouse vertical gas engine, 397 
Thermal efficiency, definition of, 578 
Thermo-chemical diagrams, 598 
Thermochemistry, 580 

Thermodynamic examination of constant pressure 
cycle, 16 

Thermodynamic examination of constant volume 
cycle, 11 

Thermodynamics, first law, 574 
Thermodynamics, second law, 579 
Thermodynamics, synopsis of, 563 
Towers, cooling, 301 

Two-cycle engmes, (Korting, American), 410 
Two-cycle engines (Korting, German), 334 
Trunk pistons, 243 
Types of American gas engines: 

Alberger Co., 486 
Allis-Chalmers Co., 475 
American Diesel Engine Co., 429 
De La Vergne Machine Co., 408 
Mesta Macliina Co., 473 
Olds Gas Power Co., 451 
Riverside Engine Co., 469 
Snow Steam Pump Works, 445 
Struthers-Wells Co., 490 
Tod Engine Co., 481 
Westinghouse Machine Co., 385 
Types of German gas engines: 

Deutsche Kraft-gas Gesellschaft, 345 
Gasmotoren Fabr’k Deutz, 324 
Giildner Motoren Gesellschaft, 354 
Oberursel Motorenfabrik, 361 
Soest, Louis & Co., 353 


672 


INDEX 


Types of German gas engines: 

Vereinigte Maschinen-fabrik Augsburg, etc., 
Nii rn berg, 344, 362 

Unions, brass, dimensions of, 309 

Valve gear, Allis-Chalmers engine, 480 
Valve gear, Struthers-Wells engine, 492 
Valve gear, Tod engine, 485 
Valve gear, Westinghouse engines, 393 
Valve gear parts, dimensions of, 212 
Valve gearing: 

Allowable fiber stress in, 204 
Constructive details of, 211 
Design of, 206 
Material, 204 

Valve gearing, adjustment of, 214 
Valves: 

Allowable fiber stress in, 194 
Constructive details of, 199 
Designs of, 196 
Kinds, consideration of, 194 
Material, 194 

Valves, governing, Alberger engine, 490 
Valves, governing, Tod engine, 486 


Valve levers, 214 

Valve lift, 201 

Vaporizers, 263 

Vaporizers, designs of, 265 

Volume of combustion chamber, 123 

Volumetric efficiency, definition of, 29 

Washers, centrifugal, 289 
Washers, gas, 288 
Water gas, 594 
Weight, atomic, 580 

Weights of reciprocating parts per sq.in. of piston 
face, 157 

Westinghouse horizontal engines, 386 
Westinghouse horizontal engines, tests of, 404 
Westinghouse producer-gas power plant with Loomis- 
Pettibone producers, test of, 401 
Westinghouse vertical engines, 385 
V estinghouse vertical engines, instructions to 
engineer, 629 

Westinghouse vertical gas engine, test of, 397 
Wrist pin: 

Diameter, determination of, 153 
Length, determination of, 153 
Wrought iron pipe, dimensions of, 309 



D. VAN NOS i'RAND COMPANY’S 

LIST OF BOOKS 

ON 

PRODUCER GAS 

AND 

Gas, Gasoline and Oil 

ENGINES 


ALLEN, HORACE. How to Design a Gas Engine. With full working draw¬ 
ings for a 7 B. H. P. gas engine. 20 illustrations. 4 * 0 . cloth. 40 pp. 
Manchester, 1907 . ne C $100 

-Modern Power Gas Producer Practice and Applications. A practical 

treatise dealing with the gasification of various classes of fuels by the 
pressure and suction systems of producer. 136 illustrations. i 2 ino. 
cloth. 334 pp. London, 1908 . net, $2.50 

-Gas and Oil Engines. A treatise on the design, construction and work¬ 
ing of internal combustion engines. 240 illustrations. 8 vo. cloth. 
548 pp. Manchester, 1907 . ne ^> $4.50 

AUDEL’S Gas Engine Manual. A practical treatise on the theory and man¬ 

agement of gas, gasolene, and oil engines. 156 illustrations. 8 vo. cloth. 
465 pp. New York, 1908 . $2.00 

BALE. M. P. Gas and Oil Engine Management. A practical guide for users 
and attendants. Second Edition. Illustrated. 121110 . cloth. 142 pp. 
London, 1907 . net > $B50 

E0TT0NE, S. R. Magnetos for Automobilists: How Made and How Used. 

A handbook of practical instruction in the manufacture and adaptation 
of the magneto to the needs of the motorist. Second Edition, enlarged. 
52 illustrations i 2 mo. cloth. 118 pp. London, 1909 . net, $1.00 






2 


D. VAN NOSTRAND COMPANY ^ LIST OF BOOKS 


BREWER, R. W. A. The Motor Car. A practical manual with notes on 
the internal combustion engine and its fuel, for the use of students and 
motor car owners. 55 illustrations. i2mo. cloth. 24s pp. London 
I9 °9- net, $2.00 

CONTENTS: History of the Internal Combustion Engine. Two-Cycle Engines. 
Large Engines. Scavenging. Mechanical Details of Construction. Details of 
Moving Parts. Stratification. Thermal Efficiency. Cause and Effect of Heat 
Losses. Testing of Gas Engines and Calculations of Results. Gas Producers. Oil 
Engines. The Gasolene Engine as Employed in the Motor Car. Power and 
Weight of Gasolene Engines. Friction and Lubricating of Engines. Clutches and 
Change-Speed Gears. Transmission Gear. Live Axles and Chain Drive. Brakes. 
The DiffcrenBa 1 Gear. Frames. Suspension. Steering. Radiation. Ignition 
Mechanism. Carburation. Liquid Fuel. Fuel Mixtures. Alcohol. Carburetters 
and the Flow of Fuel. Modernizing a Motor Car. The Care of the Car 


BROOKES, I. E. The Practical Gas and Oil Engine Handbook. Illustrated 

iomo. cloth. 192 pp. Chicago, 1905. $100 

BUTLER EDWARD^ Carburettors, Vaporizers, and Distributing Valves Used 
in Internal Combustion Engines. 100 illustrations. 8vo cloth 187 nn 

London, 1909. ‘ &00 

CA Thei^ T Th^orv‘ r” T 1 EIEDER * CRS - H - Internal Combustion Engines; 
Their Theory, Construction, and Operation. Second Edition, corrected 
Illustrated. 8vo. cloth. 611 pp. New York, 1908. net, $5.00 

CONTENTS: Introduction, Definitions, and Classification, Indicated and Brake 
Horse-Power. Thermodynamics of the Gas Engine. Theoretical Comparison of 
Various Types of Internal Combustion Engines. Various Events of the Constant- 
Volume and Constant-Pressure Cycle as Modified by Practical Conditions The 

FnZr°t agram AppHed t0 the GaS Engine ‘ Combustion. Gas 
Engine Fuels the Solid Fuels, Gas Producers. Liquid Fuels: Carbureters and 

aponzers. The Gas Fuels. The Fuel Mixture-Explosibility, Pressure, and 

Enffine? r‘ °{ ^ Engi " e - Modern T ™ es of Internal Combustion 

Engines. Gas Engine Auxiliaries: Ignition, Mufflers, and Starting Apparatus 
Regulation of Internal Combustion Engines. The Estimation of Power of Gas 
Engines, Methods of Testing Internal Combustion Engines. The Performance 
of Gas Engines and Gas Producers. Cost of Installation and of Operation 


CLERK, DUGALD. The Gas, Petrol, and Oil Engine. Vol I Thermo- 

SketT CS ^L th l^ Pet ; 0l ^ nd ° n Engine «r Wlth Historkal 
cloth Edltwn - 121 ^lustrations. 5 plates. 8vo. 

cloth. 380 P p. Lew York 1909. net> $4 00 




ON PRODUCER GAS AND GAS AND OIL'ENGINES. 


3 


CLERK, DTJGALD. The Theory of the Gas Engine. Third Edition. With 
additional matter edited by F. F. Idell. Illustrated. i 6 mo. boards. 
180 pp. (Van Nostrand’s Science Series, No. 62 .) New York, 
1903 . .50 

Dictionary. Illustrated Technical Dictionary. In six languages—English, 
German, French, Russian, Italian and Spanish. Edited according to the 
Deinhardt and Schlomann method by Alfred Schlomann. Vol. IV. In¬ 
ternal Combustion Engines. Compiled by Karl Schikore. 1,000 illus¬ 
trations. Numerous formulas. i 6 mo. cloth, 628 pp. New York, 
1908 . net, $3.00 

DONKIN, B. Textbook on Gas and Oil Engines. Fourth Edition , revised and 

enlarged. 165 illustrations. 8 vo. cloth. 568 pp. London, 1905 . 

net, $7.50 

D0WS0N, J., and LARTER, A. T. Producer Gas. Second Edition. Illus¬ 
trated. 8 vo. cloth. 320 pp. London, 1907 . net, $3.00 

G0LDINGHAM, A. H. The Gas Engine in Principle and Practice. Consist¬ 
ing of articles from “Gas Power.” 104 illustrations. 8 vo. cloth. 195 pp. 
St. Joseph, Mich., 1907 . $1.50 

-Design and Construction of Oil Engines. With directions for erecting, 

testing, installing, running and repairing. Third Edition, revised and 
partly rewritten. Illustrated. i 2 mo. cloth. 280 pp. N. Y., 1910 . $2.50 

GROVER, F. Modern Gas and Oil Engines. An exhaustive treatise. Fourth 
Edition. Illustrated. i 2 mo. cloth. 372 pp. London, 1906 . net, $2.00 

GtiLDNER, H. The Design and Construction of Internal Combustion Engines. 

Translated and revised, with additions on American Engines, by H. 
Diederichs. A handbook for designers and builders of gas and oil engines. 
728 illustrations. 36 folding plates. 4 to. cloth. 690 pp. New York, 
1910. net, $10.00 

CONTENTS: Various Methods of Operating Gas Engines.and the Gas Engine 
Cycles. General Considerations. The Various Cycles of Operation. Critical 
Examination of the Various Cyclic Events. The Design and Construction of 
Internal Combustion Engines. Fundamental Considerations. Determination of 
Principal Dimensions. General Engine Parts. Special Parts for Gas and Oil 
Engines. Auxiliaries. Construction, Erection and Tests of Modern Internal 
Combustion Engines. Stationary Engines. Portable and Self-Propelled Engines. 
The Gas Engine Fuels and Combustion in Gas Engines. Fuel Gases. Liquid 
Fuels. Fuel Mixtures. Combustion in Gas Engines. Appendix. Synopsis of 
Thermodynamics. Fundamental Principles of Thermochemistry. Some Details 
from Practice. 




4 


D. VAX NOSTRAND COMPANY'S LIST OF BOOKS 


HAWLEY, T. H. Motor Ignition Appliances. A practical treatise on the 
application of electricity in the production of the ignition spark in petrol 
motors. Second Edition. 34 illustrations. 8vo. cloth. 153 pp. Lon¬ 
don, 1906. $1.50 

HIBBERT, W. Electric Ignition for Motor Vehicles. Second Edition, revised. 
62 illustrations. i6mo. cloth. 128 pp. London, 1908. net, .50 

HISCOX, G. D. Gas, Gasolene, and Oil Engines. Eighteenth Edition, entirely 
rewritten. 412 illustrations. 8vo. cloth. 485 pp. N. Y., 1910. net, $2.50 

HOGLE, W. M. Internal Combustion Engines. A reference book for de¬ 
signers, operators, engineers and students. 106 illustrations. 121110. 
cloth. 250 pp. New York, 1909. * net, $3.00 

HUTTON, F. R. The Gas Engine. A treatise on the internal combustion 
engine, using gas, gasolene, kerosene, alcohol or other hydrocarbon as 
a source of energy. Third Edition, revised. Illustrated. 8vo. cloth. 
562 pp. New York, 1908. $5.00 

JONES, FORREST R. The Gas Engine. 142 illustrations. 8vo. cloth. 447 

pp. New York, 1909. $4.00 

JTJNGE, F. E. Gas Power. A study of the evolution of gas power, the 
design and construction of large gas engines in Europe, the application 
of gas power to various industries, and the rational utilization of low 
grade fuels. Illustrated. 8vo. cloth. 548 pp. N. Y., 1908. net, $5.00 

KENNEDY, RANKIN. Modern Engines and Power Generators. A practical 

work on prime movers and the transmission of power. Fully illustrated. 
4to. cloth. London, 1907. Six volumes. $15.00 

Single volumes. $3.00 

CONTENTS: The Prime Mover: Its Sources of Energy, Heat, Electricity. The 
Working Substance in Heat Engines: Air, Steam, Water. The Engines—Turbines; 
Air, Water, Steam, Gas. Reciprocating Engines: Single Acting, Double Acting, 
Compound, Triple, and Quadruple; Horizontal, Vertical. Engines, Rotary. Im¬ 
portant Parts of Engines: Valves and Expansion Gear, Governors, Condensers, 
Pumps, Air and Water, Bearings and Rod Ends, Lubricators. Generators—Steam 
Boilers: Cylindrical, Tubular, Flash, Water Tube, Economizers, Superheaters, Feed 
Pumps, Injectors, Ejectors, Water Supply and Coolers, Care of and Mechanical 



ON PRODUCER GAS AND GAS AND OIL ENGINES. 


5 


Stokers. Gas Generators—Coal Gas and Blast Furnace Gases: Water Gas, Mond 
Gas, Dowson Gas. Oil Fuels: Heavy Oils, Light Oils. Spirit, Gasoline, Benzoline, 
Naphtha, and Alcohol. Electric Engines—Dynamos : Batteries, Generators, Motors. 
Prime Movers Special, as made by Leading Engineers—For Mills, Factories, 
Works, etc., on Land—Turbines, Steam, Water and Air. Reciprocating, Steam, 
Water, and Air. For Marine Propulsion, Steam, Electric, Water, and Air. For 
Motor Cars: Steam, Oil, Electric. For Railways and Street Railways. Power 
Transmission and Transmitting Gearing, Belts, Ropes, Wheels, Compressed Air, 
Hydraulic Pressure, Electricity. 

LATTA, NISBET. Handbook of American Gas Engineering Practice. With 
numerous diagrams and figures. 8vo. cloth. 460 pp. New York, 
1907. net, $ 4.50 

-American Producer Gas Practice and Industrial Gas Engineering. A 

new and original work. 246 illustrations. Large 8vo. cloth. 550 pp. 
New York, 1910. net, $6.00 

CONTENTS: Producer Operation: The Producer. Cleaning the Gas. Works 
Details: Insurance Requirements. Gas Asphyxiation. Producer Types: Down 
Draft Producers. Down Draft Apparatus. The Wood System. The Tait System. 
Operation of Tait Producer. Loomis-Pettibone System. Westinghouse Double 
Zone. Westinghouse Bituminous Gas Producer. The Morgan Producer. The 
Herrick Producer. Smith Lignite Producer. Lignite Suction Producer. Wood- 
Fuel Suction Producer. Powder Fuel Producer. Marconet Powdered Fuel Pro¬ 
ducer. Moving Gases: High Pressure Plowers. Data on Moving Air. Tables. 
Solid Fuels: Ultimate Analysis of Fuels. Analysis of Gas Coal. Analysis of Ash. 
Coal Analysis. Data on Fuel Analysis. Tables. Physical Properties of Gases: 
Ignition Temperature. Chemical Properties of Gases: Combustible Gases. Diluent 
Gases. Gas Analysis: Morehead Apparatus. Gas Power: Comparison of Steam 
and Gas Power. Gas Engines: Ignition Starting. Cylinder Dimensions. Lubrica¬ 
tion. Tests. Thermal Efficiency. Industrial Gas Applications. Furnaces and 
Kilns: Brick and Tile Manufacture. Youngren Kilns. Burning Lime and Cement: 
Types of Lime Kilns. Eldred Process of Cement Clinkering. Preheating Air. 
Doherty Combustion Economizer. Combustion in Furnaces: Various Effects of 
Air Excess. Temperature, Radiation and Conduction: Miscellaneous Temperature 
Data. Heat Measurement: Pyrometry and Calorimetry. Pipes, Flues and Chim¬ 
neys: Natural Gas Measurement. Materials: Miscellaneous Data. Useful Tables. 
Glossary. 

LEWES, V. B. Liquid and Gaseous Fuels and the Part They Play in Modern 
Power Production. Illustrated. 8vo. cloth. 334 pp. (Van Nostrand’s 
Westminster Series.) New York, 1907. net, $ 2.00 

LIECKFELD, G. Oil Motors; Their Development, Construction and Manage¬ 
ment. A handbook for engineers, owners, attendants and all interested 
in engines using liquid fuel. 306 illustrations. 8vo. cloth. 287 pp. 
London and Philadelphia, 1908. $ 4.50 



6 


D. VAN NOSTRAND COMPANY’S LIST OF BOOKS 


J 


LIECKFELD, G. A Practical Handbook on the Care and Management of 
Gas Engines. Authorized translation by G. Richmond. Illustrated. 
i2mo. cloth. 117 pp. New York, 1906. $ 1.00 

LONGANECKER, E. W. The Practical Gas Engineer. A manual of practical 

gas and gasolene engine knowledge. Seventh Edition. Illustrated. 
i2mo. cloth. Cincinnati, O., 1901. $ 1.00 

LUCKE, C. E. Gas Engine Design. 145 illustrations. 8vo. cloth. 254 pp. 
New York, 1907. net, $3.00 

The work is divided into three parts. The first, treating of power, efficiency, 
and economy, gives the material, necessary for deciding on the necessary piston 
displacement for any specified output for any kind of gas, and enables the designer 
to approximately predict economy. The second part contains the data and method 
for determining the stresses in the parts and the number and arrangement of 
cylinders necessary for balance or turning effort to meet the specifications. The 
last is entirely concerned with the dimensions of the parts to resist the stresses, 
both by theoretic analysis and by empirical formulae, showing between what limits 
every principal dimension should lie. 

MARKS, L. S., and WYER, S. S. Gas and Oil Engines and Gas Producers. 

Illustrated. 8vo. cloth. 137 pp. Chicago, 1908. $1.00 

MATH 0 T, R. E. Gas Engines and Producer Gas Plants. Translated from 
the French by W. B. Kaempffert. Illustrated. 8vo. cloth. 314 pp. 
New York, 1906. $ 2.50 

MEHRTENS, A. C. Gas Engine Theory and Design. Illustrated. 121110. 

cloth. 261 pp. New York, 1909. $ 2.50 

MOSS, S. A. Elements of Gas Engine Design. Illustrated. i6mo. boards. 
l 97 PP- (Van Nostrand’s Science Series, No. 121.) N. Y., 1906. .50 

0 CONNOR, HENRY. Petrol Air-Gas. A practical handbook on the installa¬ 
tion and working of air-gas lighting systems for country houses. Illus¬ 
trated. i2mo. cloth. 75 pp. London, 1909. net, .75 

PARSELL, H. V. A., and WEED, A. J. Gas Engine Construction. Third 
Edition, revised and enlarged. 145 illustrations. 8vo. cloth. 300 pp. 
New York, 1906. ' $ 2.50 

POOLE, CECIL P. The Gas Engine. Illustrated. 8vo. cloth. 103 pp. New 
York, 1909. net, $ 1.00 



ON PRODUCER GAS AND GAS AND OIL ENGINES. 


7 


POPPLEWELL, W. C. 

Illustrated. i2mo. 


An Elementary Treatise on Heat and Heat Engines. 

cloth. 382 pp. Manchester, 1897. $2.50 


Questions and Answers from “The Gas Engine Magazine.” 

iomo. cloth. 280 pp. Cincinnati, O., 1907. 


Illustrated. 

$ 1.00 


ROBERTS, E. W. The Gas Engine Handbook. A manual of useful infor- 
mation for the designer and engineer. Sixth Edition, revised and en- 
larged. Illustrated. 321110. leather. Cincinnati, O., 1903. $1.50 


-Gas Engines and Their Troubles. With additional chapters on Design, 
Construction and Propulsion of Launches. Illustrated. 121110. cloth 
1 5 1 PP- New York, 1905. $ 1.50 


ROBINSON, W. Gas and Petroleum Engines. A manual for students and 
engineers. Second Edition, rewritten. Two volumes. 471 illustrations. 
8vo. cloth. 940 pp. New York, 1902. $8.50 

ROBSON, P. W. Power Gas Producers; Their Design and Application. 105 
illustrations. 8vo. cloth. 254 pp. London, 1908. net, $3.00 


ROOTS, JAMES D. The Cycles of Gas and Oil Engines. 43 illustrations. 
Large 8vo. cloth. 87 pp. London, 1899. $2.00 


SEXTON, A. H. Producer Gas. A sketch of the properties, manufacture and 
uses of gaseous fuels. Illustrated. 8vo. cloth. 228 pp. Manchester, 
I 9°5- net, $4.00 


SHARP, ARCHIBALD. Balancing of Engines; Steam, Gas and Petrol. An 

elementary textbook, using principally graphic methods. Illustrated. 
8vo. cloth. 223 pp. London, 1909. net, $1.75 

SIMMANCE, J. F. Calorimetry of Producer and Illuminating Gases. With 
special reference to future legislation. i6mo. cloth. 30 pp. London. $1.00 


SMITH, C. ALFRED. Suction Gas Plants. 55 illustrations. i2mo. cloth. 
205 pp. London, 1909. net, $2.00 

S0REL, ERNEST. Carbureting and Combustion in Alcohol Engines. Trans¬ 
lated from the French by S. M. Woodward and J. Preston. Illustrated. 
i2mo. cloth. 269 pp. New York, 1907. $3.00 




8 


D. VAN NOSTRAND COMPANY’S LIST OF BOOKS 


SPOONER, H. J. Notes on, and Drawings of, a Four Cylinder Petrol Engine. 

Arranged for use in technical and engineering schools, n plates 
15 x loft inches, boards. 26 pp. London, 1908. >75 

STODOLA, A. Steam Turbines. With an appendix on Gas Turbines and 
the Future of Heat Engines. Authorized translation from the German 
by Louis C. Lowenstein. Sedond Edition. 243 illustrations. 8vo. 
cloth. 488 pp. New York, 1906. ne L $ 5-00 

T 00 KEY, W. A. The Gas Engine Manual. A practical handbook of gas 
engine construction and management. Fully illustrated. 8vo. cloth. 
186 pp. London, 1908. . $ 1.50 

WARWICK, P. B. The Gas Engine; How It Works and How It Is Used. 

New Edition , revised by L. M. Schmidt. Illustrated. i6mo. cloth. 
68 pp. Lynn, Mass., 1907. 75 

WILLIAMS, E. J. Gasoline Engine Ignition. Illustrated. i6mo. cloth. 

97 PP- Cincinnati, O., 1906. $ 1 . 00 - 

WIMPERIS, H. E. The Internal Combustion Engine. Being a text¬ 
book on gas, oil and petrol engines for the use of students and engineers. 

114 illustrations. 8vo. cloth. 339 PP- London, 1908. net, $ 3.00 

WYER, S. S. A Treatise on Producer Gas and Gas Producers. Second Edi¬ 
tion. 8vo. cloth. 310 pp. New York, 1907. net, $ 4.00 

-Catechism on Producer Gas. Illustrated. i2mo. cloth. 46 pp. New 

York, 1906. net - $ 1-00 


Any Book in this List will be sent Postpaid to Any Address in the 
World on Receipt of Price , by * 

D. VAN NOSTRAND COMPANY 

Publishers and Booksellers 
23 Murray and 27 Warren Streets 'NEW YORK 












©0 




\0 c> 

-O ■ r "X \> ' 9 

S S'- . J ^ A, 

*v "* ’• *° 0° * 

* o s 0 oV * 

W* .*$&• . 

A \x^ ^ -* v ^ Ip* ° ^ c 


\°^o 


^ ftr 

*>T* C> 


IP N w * 0 

x >“ 

^'*- *- ^ 

,,p '^. c /|^lf\|r » \V V 'ki ", 

~- ' ‘l* S li' ^ ^ \ W ’ -> r £> 

' *<\.«..V‘**" a ^ .-•.% '**'V 

S * r-&>c * O (j v V r/? 7 ^ -\ • v ft 

t evV \Oa / ^ jit' , /V*^* «P ~~3T \ \ H, 

’oo^ r/Avr- : - - ^ V s 


• -V*;s.* ,>>*-;>; < *o;v* - **>\*..,>; 

V s ■ xt r *2 s* + v 



,0 



^ a 


\ V '"•, " >'WX^ *> A 

- '•«• 0 * ,'• 
V / 


<=>” *>... «X n -"/°!^,v 

. ' Jr{ - * ^ A 

*'—< ^ v 4 


•*> Vi” *<> ,> 



V 

A </> 



.0 c 


% <£' 


& 

'°*u *" * » 1 ' 4 \* s . . 

'. ^ V X V s 

o 'W' v, * 

° *$> ,p 

* ' * > * 

CA> c «« \KT ^ 

v * ^ , 
v \C> <r ^ o * v * \ \ ^ 

• 0 ^ 1 8 * % C 0 N c * Y b, 

■■ ;/- v + • 

-f. ✓ ^\Vvv^ > 

rjs y t ^ 



9 

o5 ’**■ *' 

i > ^ - y 

*m”' .^ v c< b> * ■> N o ? ,jCr 

V> s'** ^ *> 00 

s - ,^- r 


** <* ^ ^ 



«- A' ,>> ■> 

a 

”■'*/ 0 K t ^ *'• 

d* ,* 


^: x» o,. ; 

■ 

/> 

/* c* 

* ^ 

° ^r, \V 

° z ^ 4 

s- 

.. . « v/ 

■i . ' 4* '.^K s# % a' 

0 « V ^ K , •©, '/ 0 . s ^ J& , . , <. y o . 

^ ■/ " 






o 


W 






x. N ^ ^ ' ,. - 

^ ^ ^ .Po^\o° 

* * » > K A> „ 1 * 0 


<V V 

o o x 


\ ,V-' 

" " \^ s’*/ 

V. ^ ; 


0 


/ -v - 


vV 



>v 


^ % 


■■^ ^ °,, / ? J; ^ y s « vl " ^ 

^ " .- ^ ' v jedlTfei . 1 + ‘ ^ 

L *v-v 




b ^ ^ ,0 







p 1 ’ &*U V 

r © ^v‘ .O'- 

J ‘ * r b * > 1 

° K - 

% v" 


vNJ\ - 7 paw. rrfTTj* 7 . ’ 7 

-v .c^ o 

o^‘ * 

■' ,/ ,ti., <<. 

0 '+ SO' 

,■■■: ^ # j 

, o \0 o_ 

^ *n° ^ 0 ' - 

r i f > O’ * ^ * 0 /* 'C* V' s ' ' // > o 

^ (A at jf ^rA, r V oc, ^ o ^ ^ 

ri - ^c> « A { ■': '., 0 V?n 


^ - .— ■ ™ 

» __ _ 

t r - 0 

o> -7% - W^ v ° 9 

\v” VL-. . > ^ 

%/•' ;. > e / 

V ^ ♦ 1 

xV * * -b 

\ $ * mmJ/ >L \ * «y ; ; -L j - : .: %$ . 

; ..^ \ %WMW+ *, ", 

a° vi« y °^ * A * 

py? « v 1 8 H a\ 0 N G , ^ 

G>-s r VJ v -/ ^>. JTO C G-'li 

, ° 0 v v 1 f ° 

'- '“o* : §mw^'. *'- $ :*v 

V®sP c .f - . 

^ ^ ^ N a: c^ ^ JJ r*» i. 

t 5 N° ^ * tf . \ 0 \^ V 

x/,*;- 0 '. ^ v\,< 

\/ * £%&• 

^ ^ ^ z ^ 


9 AV 

x> N 


?y 0c ^ 

^ \ 

^*. ^ ^ n 

•••y »<■•/% "'-V •.-■, 

% *» v :MK\ b ^ - 

A ^ o 

'bf. M/ '«■ * 




* «V ^ Lr <^'*^ xr '^' '^' n#K 0 * -.V* ^ b’K/ tf 

A' <“ y 0 * V ■*■ A Ol ' >/ <'z s ^ S * ^ o 

0^ v *v.«, A c 0N Cjf -Z^ 7 **' 0 ^ X.», > 0 

^ it , • ,>^ v a _ c a iiTt , <v , c ^ MElW/syZ* 





' J bo x 




Py/ *n'*' v \^ s , () -To °* 



O 

Ac. t " J ~f. 

\ .- A *v o' ^ 

o *, s a* ^ u 

p Y tn ^ * •.' * \\ % s ,,, ^ * 

^ ^ aTx^^/U ^ 4V. ^ ft 

0 Vj» ® 



^ ' 

Ci. 

( * * N 

\ 

% * 

« 1 ' 1 0 

* 

\* 


i® 

? «B 


^ %W-.' 


* V 9 

' / .* 

V*®-. ‘ 

+ 

<?• I- 

0 r x • « ^ 

T ^ v< \ : ’ 

^ 0 * X * ,\ +‘ * 

1 \ ^ c « N 0 A v b^ ” ,.o" - 1 1 " 

^ 4 • 0 ^ v 

^ ■* ^ y 


> .— 

. ; #% vr r •»■ 

a \ . v- 

a\ 0 N G . ^ 7 7 ♦ * S S v 

c w * r\ rv‘ « ' 






0 l 

ft "<> 

A A 

T y 0 9 


<$> 

-P 

/ 

& *y 

- 


\° 



9 

c 
>> 

k ft _0 O /u 

^ * .0 N 0 ? ^c * . ,A*’ . 

A 0 V ^ ' *«A o V v '' 

■b- & »w A v V a. * 

: v .. ;■' ,■; ■, = ^ ^ ^ 

• v 

• ^ ° 4. ,3-XUr ± V 


<0 , , <b y o.jA A '^'''J N s' i0^ <* 

°. ^ P 0 ^ v’ % % ^ A N ''A V c . 0 > . v > • « 





/V ^ ,‘J’ 9- ’ “ ° /■ ”c 

s •- % .0 .w*; b- 4- 

- *< S as ' >/ cP *\ V 

V ^ z - - * 



■’o 0^ 


n 0 ’ .o'* c c ^ A- 

A- ^ ^ o z % » 1 ’ vVs'"' 


/ ' 7 ^- 









^7:Ca s ^ 

o r\» <« v 



Tp, o 

TV> ^ 

* 

• * ^ y 0 * 4 a' c 0 N c « 

c“ v^ • 

- b, a' 5 9 


» 

4? 

a 

^ A 




* .0 r* 0 ’ s. 0 



^ A' 

' 1 * * s ' A , 11 » 

-* C 0° .V s 

✓ ^ V 

o o N ♦ 

i> 

o5 

v ^ 

•v _^* ^ y 




o, x ^ 4 A >c O b, 

^ c 0 N c « ' 7 g 

. ^ O 

A^ <*. cp y* 

- V s c , T 

ft * 

^ 0 Oft - 

^ ^ IP p a - * **z/yjy& * ■ 

C!> * ft s. 0 ' 'O’ 




.-. ;/<3r ,> ^ 

^ \ 


L. ft « 


*0 



0 




u * 

p ^ ‘ 

V s %■*'**■' 

A' * 

A © v^. C^ 

x> %> </;, 

'V •■>,. 

s 





^ * * * s s \\ Lift <J / 0 * y <y < ** 

A^ • v * 1 * M / O rv> 0 N G ^ 

A tf/fTfe! + ° 0 , 

' • " "O O' c \. - * ••- ; v : ", ^ ^ 

rv ^ * VV 

A -A. 

1 W <^ V V * * , ^ N 0 ’ ^ '*...* ' -•O'" 'O 


*» r °./. * 'm o ’ x# C © * o’ W 

'•"'% fc v v:*. > >^aV 
-- ^ .*/8Sfcv« %/ *^‘ v 

<\V «/» 

A \> »/>. c 

V *■>*. . J V -^ 7 <c\ \|- - u.' xv» 


■V 

A. 



> 





^ %l; \V>se s, \V 


✓ a. 4 

1 » D s o •> ^ 

%> 

C A V N % ' 

a f . 

VP V, 

•/\ v\ V ^ V' 




.V s A 


■i. A 

.> rW‘ / ^ . 

< x vl , 'om* A </^; v \x'.A 

,* l "< 'o <V , o N C 7 - ■> S a\ 

X* *\r-/r??^i ° 0° * C ^Nx % ^ 

A 4 .^ *4 R .-. *■ • ■' A*^. ' Kp -1 

^ K 

X° °* * 

* , , , . ''\o° C o '' 

\L r *% 

.. .XX- - -x ... _ JP Sjtifoi'. * 




'V 




A \ . \ i B f. ^{- 

* ■>- -*« * . o> JL 

I 


: '.f®: 

* a 0> N <U * 7 ** s ' X A "?J ^ 

(V t 0 '• <■ % A V . »• ' 1 * V 0 C 6 N ° * '<8. * * ' 

^ V ° 0°/ t ^ **% J 

. . , „' 7 V "»»* ,= vi^'. ^ v* 

r •»>* x°°<.. *,v~h-■ »-; 

,* A A '. -- ^'- a o - c c w.' A % x 

"'* ' »’••/ -> '••'*> s '” ,> v .... A... 



^' V rxO 

« ' ' 6 A. v «« . * ■> N o ^ ^ ‘in 4 ’ J u ' "6 " 9 s 0 ■’ 

v> A'"A "> A 0^ s' 7 ,, v> A »'• 

■** ' *t»*\ x. A .>VA' ^ / *'€%»'■ A A .' .«■> 

$ y *r ° WMW * A* - ' '^'i tWr ° C ^ ^-W " 

l '• " - ' ^ -v < fi 0 'V «>>. ^ v v^ n \K ^ ■c >» w 

. V* ✓ ^ -a \L' > ^6 

" //7 s' .,, % 'o> A C' C A 

> N . * ., .0' 0 c « , * 3 ^ 


A x. % 

'A 7 '’ * ' sX'. , i ' « « “V-- 

% ■% i ■ *XrvL'. ° 

A' JTvl/ , o2 L, 


<* 


tP 

v> ^ 



^ A 

► V aO^ 

cf s"''r* *c 

v x ^ -P 








•" ^ & *Z*. + %' i) 4 r± ^ r™ 

r V'*in‘ X / s^/^'sTo’* ^ 
// c- \* s-' °^ ^ ,0 C s' ca v 



<a /- * ■" *V ,\0 

^ Nn 4 4 ° 
s *o, * 1 cy> x s ‘ 

j~i i , « >\ ’ \ 








